Method for determining and communicating a vehicle dynamics limit range of a steering system and steering

Abstract

Method for determining and communicating a driving dynamics limit range of a steering system (1), where a nonlinear kinematic tire trail is instructed (n R ) is determined, where based on the determined nonlinear kinematic tire trailing (n R ) i) feedback (21) is generated at a steering handle (51) of the steering (1) and / or ii) an assistance system function (55) is activated; where to determine the nonlinear kinematic tire trail value (n R ): - based on the measured and / or estimated lateral acceleration (ÿ v ) a tire lateral force (F yv ) is determined, - starting from the nonlinear kinematic tire trail value (n R ), a constructive tire trail value (n K ), a track lever ratio (i kin ) and the specific tire lateral force (F yv) a first rack force (F Zst,1 ) is determined, - starting from a measured engine torque (M mot ) an electric machine (8) used to apply a steering torque a second rack force (F Zst,2 ) is determined, - the specific first rack force (F Zst,1 ) and the specific second rack force (F Zst,2 ) are compared with each other, and - based on a comparison result, the specific tire lateral force (F) yv ) is corrected, whereby the nonlinear kinematic tire trail value (n R ) starting from the corrected tire lateral force (F yv,corr ) is determined.

Description

[0001] The invention relates to a method for determining and communicating a driving dynamics limit range of a steering system and a corresponding steering system.

[0002] In automotive engineering, there is great interest in knowing the operating point of a tire, especially at its limits. This information helps, for example, to prevent the tire from being driven over, ensuring that the tire operates at its optimal point.

[0003] DE 10 2009 006 211 A1 describes a vehicle dynamics control system for a two-track motor vehicle with a steering system at the front wheels. This system monitors the vehicle's approach to its lateral dynamics limit and, if necessary, initiates a suitable action. This action either alerts the driver to the approach to the limit or modifies the vehicle's driving state to increase the distance to the limit. The approach to the lateral dynamics limit is derived from an evaluation of the curve of the restoring torque of the wheels exhibiting at least a slight steering angle, or a substitute parameter for this torque, as a function of the steering angle.A restoring torque that does not increase further or decreases with increasing steering angle or slip angle, or a corresponding equivalent parameter, is used as an indicator of reaching nonlinear driving behavior and thus the lateral dynamic limit. The restoring torque can be measured or appropriately estimated, preferably as the aforementioned equivalent parameter in the form of a measured or estimated force on the tie rods of the steering system that cause a steering input at the steerable wheels.

[0004] DE 10 2013 009 399 A1 describes a method for detecting a critical driving situation of a vehicle with the following steps: detecting a deviation from a predefined relationship between the rack force of a rack and the steering angle of the vehicle, and, in response to the detected deviation from the predefined relationship, detecting a critical driving situation of the vehicle.

[0005] DE 10 2017 213 535 A1 describes methods and systems for determining vehicle skid conditions, including conditions that indicate skid before the vehicle actually skids. The methods and systems receive motion parameters of a vehicle based on acquired signals from at least one vehicle sensor, an electronic power steering system, and an inertial measurement unit. The methods and systems estimate the tire caster based on the rate of change of the self-aligning moment with respect to the axle lateral forces. The methods and systems determine the skid conditions based on the estimated tire caster. The methods and systems control at least one vehicle feature in response to the determined skid conditions.

[0006] DE 102022 205 552 A1 describes a method for generating haptic feedback force by adjusting a hand torque on a steering input device of a motor vehicle's steering system. In this method, a first and a second rack force value are obtained from two different models. The rack force values ​​from both models are processed to form a quotient that determines a feedback force value. This is achieved by modifying the second rack force value by the quotient, and the feedback force value is then transmitted to a steering reference torque generator of the steering system, which adjusts the hand torque on the steering input device as a function of the feedback force value.

[0007] German patent DE 10 2022 119 666 A1 describes systems and methods for determining whether a vehicle is in an understeering or oversteering situation. The system comprises a control circuit coupled to an IMU and an EPS and programmed to: calculate an axle-based pneumatic track for a steered first axle using IMU measurements and EPS signals, and estimate a saturation level as a function of the distance between the axle-based pneumatic track and zero. For an unsteered second axle, the system estimates an axle lateral force curve with respect to a slip angle of the second axle and a degree of saturation depending on when the axle lateral force curve transitions from positive to negative values ​​with respect to the slip angle. The saturation levels of the first and second axles are integrated.The system determines whether the vehicle is in an understeering or oversteering situation, depending on the integrated saturation values.

[0008] The invention is based on the objective of providing a method for determining and communicating a driving dynamics limit range of a steering system and a steering system.

[0009] The problem is solved according to the invention by a method with the features of claim 1 and a steering system with the features of claim 9. Advantageous embodiments of the invention are set forth in the dependent claims.

[0010] In particular, a method for determining and communicating a driving dynamics limit of a steering system is provided, wherein a non-linear kinematic tire caster value is determined, in particular starting from at least one lateral acceleration detected and / or estimated at a steering axis of the steering system, wherein, based on the determined non-linear kinematic tire caster value, i) (in particular additional) feedback is generated at a steering handle of the steering system and / or ii) an assistance system function is controlled.

[0011] Furthermore, a steering system is created, in particular comprising a control device, wherein the control device is configured to determine a non-linear kinematic tire caster value, in particular starting from at least one lateral acceleration detected and / or estimated at a steering axis of the steering system, and furthermore, based on the determined non-linear kinematic tire caster value, i) initiate (in particular additional) feedback at a steering handle of the steering system and / or ii) control an assistance system function.

[0012] The method and steering system make it possible to determine and communicate a driving dynamics limit of a steering system to the driver and / or to control an assistance system function based on this limit. The communication is intended to include, in particular, (additional) feedback and / or the control of an assistance system function. One of the fundamental concepts of the invention is to use, in particular, a nonlinear kinematic tire caster value as a measure of the driving dynamics limit. This specifically exploits the fact that the value of the nonlinear kinematic tire caster changes significantly in the driving dynamics limit with respect to a slip angle. The nonlinear tire caster value is determined, in particular, based on at least one lateral acceleration detected and / or estimated at a steering axis of the steering system.Based on the specific nonlinear kinematic tire trail, (additional) feedback is generated at a steering handle and / or based on the specific nonlinear kinematic tire trail, an assistance system function is controlled.

[0013] A steering system can be an electromechanical steering system or a steer-by-wire steering system.

[0014] Controlling the assistance system function can, in particular, include triggering the assistance system function and / or an associated assistance system. For example, the assistance system can provide automated or semi-automated lateral guidance. Furthermore, it can also be provided that, based on the specific nonlinear kinematic tire tracking, a type and / or strength of support intervention by the assistance system function is determined.

[0015] Components of the steering system, particularly the control unit, can be designed individually or collectively as a combination of hardware and software, for example, as program code executed on a computing device, especially a microcontroller or microprocessor. However, it is also possible for components to be designed individually or collectively as an application-specific integrated circuit (ASIC) and / or a field-programmable gate array (FPGA) and / or a graphics processing unit (GPU) and / or a digital signal processor (DSP). The control unit can, in particular, include at least one computing device and at least one memory. Furthermore, the steering system can have a communication interface through which it can communicate with an assistance system that performs the assistance system function and / or with a vehicle.

[0016] The following details relevant to understanding the procedure described in this disclosure: The tire lateral force F acting by the vehicle on a rack of the steering system yv The force is drawn in through the tire and transmitted via the steering axle to a steering gear. Between the rack force F present at the steering gear. Zst and the tire lateral force F yv The tire lies including two lever arms: Firstly, the tire caster n ges , consisting of the constant constructive tire trail n K , which is known to manufacturers (the design caster is in particular the distance from the point where the extended steering axis intersects the road surface to the tire center), and the variable, non-linear kinematic tire caster n R and secondly the kinematic track lever ratio i kin (see Fig. 3) The kinematic track lever ratio i kinis a function of a rack position i kin = f(x Zst ), predetermined by the steering design and therefore known. The constructive tire caster n K is determined by an axle design. The nonlinear kinematic tire caster value n R depends on the tire and is also a non-linear function of the slip angle α v .

[0017] Due to the tire lateral force F yv The tire tread deforms. This creates a force that acts against this deformation. Furthermore, the geometry of the deformation results in a center of gravity (see Fig. 4), where this force acts. The distance of this center of gravity from the tire center corresponds to the tire caster n. ges .

[0018] Under small forces, the tire contact patch deforms linearly and progressively. Under larger forces, the tire slips at the rear of the contact patch, resulting in a degressive behavior. This increases the lever arm, i.e., the tire caster. R , not linearly, but exhibits degressive behavior.

[0019] Through the connection FZst=Fyv⋅(nR+nKikin) As is evident in the stationary case, the force level of the rack decreases when the limit of the tire is reached.

[0020] The tire lateral force can be determined from the lateral acceleration ÿ. For this purpose, the lateral acceleration ÿ is used. v required on the front axle. Furthermore, the wheel contact mass m v the front axle is required. If the lateral acceleration at the front axle is known, the tire lateral force can be calculated by Fyv=y¨⋅mv be determined.

[0021] From the relationships described above (especially equations 1 and 2) it follows that only one quantity is missing: the kinematic tire trail n R The procedure for determining it is described below.

[0022] For many tires, both a curve of the tire lateral force F yv above the slip angle α v , as well as a course of the nonlinear kinematic tire trail n R above the slip angle α v at least roughly known. For example, these curves can be determined in the form of characteristic curves during tire measurements under given measurement conditions (see Fig. 7 and Fig. 8).

[0023] A tire lateral force can now be calculated from a measured or estimated lateral acceleration. Subsequently, an inverse tire model can be used (see...). Fig. 7) a slip angle α vIt can be calculated from the tire lateral force. This slip angle is used to determine the nonlinear kinematic tire caster n. R to determine (e.g. starting from the exemplary characteristic curve in Fig. 8).

[0024] The following are examples of the inverse tire model and the trailing model:

[0025] The tire lateral force F yv is the input variable of the inverse tire model. In this tire model, the slip angle α is used. v due to the tire lateral force F yv determined. These two quantities are interdependent. The tire lateral force can be determined, for example, by a Pacejka model according to the regulation. Fyv=DPacejka ⋅ sin(CPacejka ⋅arctan(BPacejka⋅αv−EPacejka) ⋅(BPacejka⋅αv−arctan(BPacejka⋅αv)))) be calculated, where B Pacejka , C Pacejka , D Pacejka and E Pacejka These are the parameters of the model.

[0026] The force scaled to a value of 1 corresponds accordingly to: Fyv,scaled=sin(CPacejka ⋅arctan(BPacejka⋅αv−EPacejka) ⋅(BPacejka⋅αv−arctan(BPacejka⋅αv))))

[0027] The factor D results from the tire contact force and represents the maximum point of the Pacejka curve: D=mv⋅g=m⋅lhl⋅g where l is the wheelbase and l h The distance from the center of gravity to the rear axle. Conversely, the slip angle depends on the tire lateral force when the model is inverted (by inverting the characteristic curve). Accordingly, a slip angle can be determined from the tire lateral force. However, the characteristic curve can only be used up to the inflection point of the Pacejka curve (see Fig. 6) Otherwise, inversion leads to ambiguity. A unique characteristic curve is not possible in Fig. Figure 7 shows the (dashed part of the curve up to the inflection point). Alternatively, the tire lateral force can be determined using a characteristic curve.

[0028] Using this characteristic curve ( Fig. 7) a slip angle α is determined v determined. This slip angle α v is used to determine the nonlinear kinematic tire trail n R used. The nonlinear kinematic tire trail n R is also modeled in particular using a Pacejka model with the model parameters B, C, D and E (see Fig. 8): nR=D⋅sin(C⋅arctane(B⋅αv−E⋅(B⋅αf−arctan(B⋅αv))))

[0029] This tire caster angle is the lever arm used for calculating moments. Alternatively, the nonlinear kinematic tire caster angle can be used. R can also be determined using a characteristic curve.

[0030] How to learn from the Fig. 7 and Fig. 8 detects, means a greater tire lateral force Fyv a larger slip angle α v and thus a smaller nonlinear kinematic tire trail value n R (resulting in a smaller rack force at the end). This is precisely what causes a drop in force at the limit.

[0031] In one embodiment, the nonlinear kinematic tire trail is determined by taking into account a deviation between a quantity determined from the detected and / or estimated lateral acceleration and a quantity determined from a detected motor torque of an electric machine used to apply a steering torque. This allows system friction within the steering system to be considered. The determined quantities can, in particular, be rack forces, as in the following embodiment.

[0032] It is intended that the following will be used to determine the nonlinear kinematic tire trail value: - a tire lateral force is determined based on the measured and / or estimated lateral acceleration, - starting from the nonlinear kinematic tire caster value, a design tire caster direction, a track linkage ratio and the determined tire lateral force, a first rack force is determined, - starting from a measured motor torque of an electric machine used to apply a steering torque, a second rack force is determined, - the definite first rack force and the definite second rack force are compared with each other, and Starting from a comparison result, the specified tire lateral force is corrected, whereby the nonlinear kinematic tire caster is determined based on the corrected tire lateral force. This is repeated continuously, particularly until the first rack force converges. By comparing the specified first rack force and the specified second rack force, and subsequently correcting the specified tire lateral force, the values ​​of the specified first rack force and the specified second rack force become similar (particularly iteratively), whereby the nonlinear kinematic tire caster also changes its value.If the determined first rack force and the determined second rack force converge against each other, a reliable estimate for the nonlinear kinematic tire trail is available, which also takes into account friction within the steering system.

[0033] The preceding embodiment is based in particular on the following line of reasoning: Using the preceding equations and relationships, the force acting on the rack due to the vehicle's reaction can be determined. Furthermore, in steering systems, the engine torque M motThe torque is measured by an electric motor used to apply steering torque to the steering system. The electric motor is mechanically rigidly connected to the rack, meaning the torque must necessarily be qualitatively proportional to the rack force. The only interactions between the motor torque and the rack are friction and a finitely rigid connection, resulting in some information loss. However, a rough level and qualitative curve are the same outside the range of uncertainty caused by friction. Therefore, the motor torque can be used as a ground truth for determining the rack force.

[0034] The determined first rack force is compared with the motor torque of the electric steering machine converted to a rack plane, i.e., with the second rack force, thus with the assumed or hypothetical ground truth. If the force from the lateral acceleration (first rack force) is greater than the force from the motor torque (second rack force), then, assuming correct characteristic curves, this is due to the tire caster n. RTo explain, a comparison result—that is, a difference between the two rack forces—can be used to determine whether a larger slip angle is required than originally identified in the inverse characteristic curve. This information is then fed back in order to correct the operating point (i.e., the tire lateral force) of the tire in the characteristic curve in the next calculation step and to adjust the slip angle estimate so that the two determined rack forces are quantitatively identical. In this way, the slip angle and the nonlinear kinematic tire caster are determined or estimated with improved accuracy.

[0035] In particular, the second rack force based on motor torques is compared with the first rack force calculated from the lateral acceleration, and a correction factor c is derived from the ratio of these two forces. corrcalculated. This correction factor scales the tire lateral force F. yv, The slip angle is calculated from the lateral acceleration, which changes the operating point in the characteristic curve and thus adjusts the slip angle. This corrected slip angle is used to recalculate the nonlinear kinematic tire caster from the model. This lever arm is then also used to recalculate the first rack force from the lateral acceleration.

[0036] As part of force correction, the correction factor is particularly important ccorr=Fzst,2FZst,1 calculated. A value of the maximum force of the Pacejka curve. D=mv⋅g=m⋅lhl⋅g The correction factor is applied if the motor force and the lateral acceleration force diverge. Therefore, if the motor force is less than the lateral acceleration force, the factor D is scaled down, resulting in different values ​​on the Y-axis of the characteristic curve. Fig. Shift 6 downwards. This results in a larger slip angle being determined from the characteristic curve for the tire lateral force calculated from the lateral acceleration, due to the scaling, and consequently a smaller value for the nonlinear kinematic tire caster value n. R .

[0037] In one embodiment, the motor angular velocity of the electric machine is detected, and friction compensation is performed on the second rack force, estimated from the motor torque, based on the detected motor angular velocity. This compensates for friction in the electric machine, as explained below. When comparing the two determined rack forces, there is a margin of error due to friction: the lateral acceleration signal contains no friction, but the motor torque signal contains several hundred Newtons of system friction. Therefore, this method can only perform a slip angle correction if the rack force resulting from the motor torque is above a certain limit.This is unproblematic with a high coefficient of friction, as the nonlinear region of the tire, identified in this way, only occurs at very high forces. However, with low coefficients of friction, for example on an ice- or snow-covered road, the nonlinear region of the tire is reached at significantly lower forces. Here, the area of ​​uncertainty caused by system friction becomes problematic. For this reason, one embodiment uses friction compensation to reduce this area of ​​uncertainty. This is described below by way of example.

[0038] In the steering gear, in addition to the engine torque, a rotor angular velocity φ̇ is usually also included. Mot The friction is available as a measured quantity. The friction of the system can be estimated using this quantity and a friction model. For example, the LuGre friction model is used here.

[0039] In this model, the dynamics of friction are calculated using a bristle model. The dynamics of the bristle motion z are calculated using the following differential equation: dzdt=φ˙Mot−|φ˙Mot|⋅zg(φ˙Mot)

[0040] The function g(φ̇ Mot ) is calculated, for example, using a Stribeck model: g(φ˙Mot)=FC+(FS−FC)⋅e(−φ˙Motφ˙s)2

[0041] In this equation, F represents C the Columb friction, F S the breakaway force and φ̇ s The Stribeck velocity. The bristle deflection and the Stribeck curve are then used to calculate the frictional force. F=σ0⋅z+σ1⋅dzdt+σ2⋅φ˙Mot

[0042] Here, σ0 corresponds to a stiffness, σ1 to a damping coefficient, and σ2 to a velocity-proportional component. The friction level present in any given situation can be identified from measurement data. For example, the friction level may be higher at lower ambient temperatures. Therefore, complete elimination of friction is not guaranteed in every situation, but the range of uncertainty is reduced.

[0043] This estimated frictional force is then subtracted from the rack force calculated from the motor torque, thereby reducing the frictional force contained in the signal and thus the uncertainty range. The second rack force, corrected for friction, is then used for comparison.

[0044] In one embodiment, the lateral acceleration is detected by means of a lateral acceleration sensor arranged on or near the steering axis. This allows the lateral acceleration to be detected directly on or in the immediate vicinity of the steering axis.

[0045] In one further embodiment, the lateral acceleration sensor is arranged in or on a steering module of the steering system. This allows for a particularly simple design. In particular, this eliminates the need for complex wiring. Specifically, the lateral acceleration sensor can be arranged directly on a circuit board (PCB) of the steering module, especially the steering module's control unit.

[0046] In one embodiment, the lateral acceleration is estimated based on measurements taken by an inertial sensor. This allows the vehicle's existing inertial sensors to be used, eliminating the need for a separate lateral acceleration sensor. If the lateral acceleration at the vehicle's center of gravity is known, the lateral acceleration at the steering axis can be determined using Euler's theorem for rigid body kinematics and the displacement between the front axle and the lateral acceleration measurement point (P). x |P y |P z ), the yaw acceleration ψ̈, the pitch acceleration φ̈ and the roll acceleration κ̈ are determined [ax,Pay,Paz,P]=[axayaz]+[κ¨φ¨ψ¨]×[PxPyPz]+[κ˙φ˙ψ˙]×[[κ˙φ˙ψ˙]×[PxPyPz]] where α x , α y , α z The accelerations are at the center of gravity.

[0047] In one embodiment, a wheel contact force is detected and / or estimated, and a wheel contact mass is determined based on this detected and / or estimated wheel contact force. This determined wheel contact mass is then taken into account when determining the tire lateral force. This allows a detected and / or estimated value for the wheel contact force to be used.

[0048] In a further developed embodiment, the wheel contact force is adjusted based on detected and / or estimated suspension travel and / or accelerations. The wheel contact force m vThis can be continuously adjusted during driving by means of suspension travel and / or acceleration. When the vehicle accelerates, this results in roll, pitch, and / or yaw movements, which in turn cause a transfer of wheel load. During acceleration, the wheel load on the front axle is reduced in particular. This reduces the maximum transmissible forces. By using the acceleration information, this effect can be estimated and thus taken into account by decreasing or increasing the wheel contact force accordingly.

[0049] In one embodiment, a feedback torque for a steering handle is determined and provided based on the supplied first rack force. The first rack force F is determined from the lateral acceleration. Zst,1This is used to provide the driver with steering-induced feedback (this feedback is not the same as the feedback used to convey the vehicle's dynamic limits). For this purpose, the initial rack force is either converted into a feedback torque in the steering gear and transmitted to a force-feedback actuator on the steering wheel, or the initial rack force is transmitted directly to the force-feedback actuator, where the conversion into a feedback torque for the driver takes place. A lateral acceleration sensor on the front axle or steering system can significantly enhance the frequency content and eliminate friction from the feedback torque signal. This also provides a method for determining the feedback torque.

[0050] In one embodiment, the strength of the feedback and / or the control signal is determined based on a specific value of the nonlinear kinematic tire caster value. This allows the criticality of the vehicle's dynamic limits to be conveyed as part of the feedback. Specifically, it can be provided that the more critical the situation, i.e., especially the smaller the value of the specific nonlinear kinematic tire caster value, the stronger the feedback to the driver and / or for the assistance system function.

[0051] In one embodiment, the specific nonlinear tire caster value is normalized and taken into account. This allows for comparable feedback, particularly for different vehicles or tire characteristics, and simplifies further processing of the normalized value, as the processing can be performed in the same way for all steering systems. The normalization is specifically based on a maximum value of the nonlinear kinematic tire caster value in the corresponding characteristic curve. For example, the normalized nonlinear kinematic tire caster value can be represented and considered as a ratio: cLimit=nRnR,max where n R,max the maximum value of the nonlinear kinematic tire trail value n R is.

[0052] There are several ways to communicate the limit to the driver, which are described below using some examples.

[0053] In one embodiment, the dependence on the specific nonlinear kinematic tire trail is determined by a predefined characteristic curve. This allows for direct control over the strength of the feedback transmitted to the driver and / or the assistance system function, reflecting the vehicle dynamics limits expressed by the nonlinear kinematic tire trail. The characteristic curve is stored in a memory of the control unit and can be retrieved from it as needed.

[0054] In one embodiment, feedback and / or control is only initiated once a predefined threshold value for the specific nonlinear kinematic tire caster value is reached. This allows a threshold to be defined below which feedback to the driver and / or the assistance system function should occur. Due to the characteristics of the nonlinear kinematic tire caster value, it is specifically intended that the threshold value must be undershot for feedback or control to occur.

[0055] In one embodiment, the feedback is provided in the form of a periodic signal and / or an oscillation. This allows the driver to receive feedback that does not influence or impair the steering behavior itself. The periodic signal and / or the oscillation is particularly referred to as steering-induced feedback (in this disclosure, specifically as F). Zst,out (designated), which is provided to the driver at the steering handle, is overlaid.

[0056] The limit can be expressed, for example, as the additional oscillation F Grenz The feedback loop is represented. The amplitude F A and / or the frequency f of the oscillation force F Grenz This can depend on the ratio c GrenzThis can be represented from the maximum value of the nonlinear kinematic tire caster value at the current operating point. Here, this additional oscillation is superimposed on the feedback force: FGrenz(t,FZst)=(1−cGrenz)⋅FA(FZst)⋅sin(2⋅π⋅t⋅f(FZst))

[0057] The frequency of the oscillation f can be chosen to be constant. However, the frequency of the oscillation f can also depend on a currently applied force level F. Zst This is displayed because, in reality, the frequency of tire vibrations also depends on the force level. This provides the driver with clear information about the tire's operating condition.

[0058] The amplitude and frequency of this oscillation can also be determined by characteristic curves, which, for example, depend further on the factor c. Grenz and / or a vehicle speed v Fzg and / or a determined rack force F Zst,outcan be defined.

[0059] The total feedback (as a force) is therefore: FZst_Feedback=FZst,out+FGrenz

[0060] Based on this, a corresponding hand torque results at the steering handle.

[0061] In one embodiment, the feedback is taken into account in the form of a factor that scales steering-induced feedback at the steering handle. This allows the steering-induced feedback to be directly adjusted.

[0062] In one variant, the limiting region in the feedback loop is represented by a sudden and exaggerated drop in the feedback force. Here, too, the ratio c is used. GrenzThe maximum value and the currently determined nonlinear kinematic tire caster are taken into account. This ratio can be arbitrarily shaped by characteristic curves and used as a factor of the steering-induced feedback (i.e., in particular as a factor of the calculated rack force F). Zst,out ) are taken into account in order to, for example, create a drastic, exaggerated drop in feedback to the driver. The drop in the feedback level is determined via the rack force F. Zst,out represented, which around a characteristic curve as a function f Gain (c Grenz ) of the ratio c Grenz is factored so that the entire feedback F Zst_Feedback at the steering handle results in: FZst_Feedback=FZst,out⋅fGain(cGrenz)

[0063] In another variant, the force level at the limit can be used reciprocally, depending on the ratio of the maximum value to the currently determined nonlinear kinematic tire caster, to increase the feedback force level. This would have a stabilizing effect, as the tire's limit is left when the steering angle is reduced. The increase in force provides a recommendation to the driver to implement precisely this behavior. Increasing the force level as a function of the nonlinear kinematic tire caster, or the aforementioned ratio, can be implemented, in particular, as a suitably designed function or characteristic curve. The relationship here is fundamentally the same as in Eq. 15.

[0064] Further characteristics for the design of the steering system emerge from the description of the process configurations. The advantages of the steering system are the same in each case as in the configurations of the process itself.

[0065] The invention is explained in more detail below with reference to preferred embodiments and the figures. These show: Fig. 1 a schematic representation of an embodiment of the steering system; Fig. 2 a schematic flowchart to illustrate embodiments of the method; Fig. 3 a schematic representation to clarify the quantities mentioned in this revelation and the relationships described in the general description; Fig. 4 a schematic representation to clarify the quantities mentioned in this revelation and the relationships described in the general description; Fig. 5 a schematic representation to illustrate the quantities mentioned in this disclosure and the embodiment in which the lateral acceleration is estimated from measured values ​​of an inertial sensor; Fig. 6 an exemplary characteristic curve (tire lateral force versus slip angle) according to a Pacejka model; Fig. 7 An exemplary characteristic curve (slip angle versus tire lateral force) according to an inverted Pacejka model (inverted characteristic curve of the Fig. 6); Fig. 8 an exemplary characteristic curve (nonlinear kinematic tire trail versus slip angle) according to a Pacejka model; Fig. 9 a schematic flowchart to illustrate one embodiment of the method; Fig. 10 a schematic flowchart to illustrate another embodiment of the method.

[0066] The Fig. Figure 1 shows a schematic representation of an embodiment of the steering system 1. Fig. Figure 1 shows an example of a steer-by-wire steering system. However, the steering system 1 can also be an electromechanical steering system. The steering system 1 comprises a steering wheel module 2, a steering module 3, and a communication link 4 between the steering wheel module 2 and the steering module 3. The steering system 1 is, in particular, arranged in a vehicle 50. The steering system 1 operates as a steer-by-wire steering system in a manner known per se. The steering system 1 is configured to perform the method described in this disclosure. The method is described in more detail below with reference to the steering system 1.

[0067] The steering wheel module 2 is designed to detect a steering input from the driver at a steering handle 51 (e.g., steering wheel) and, since this example involves a steer-by-wire system, to generate a feedback torque 20 at the steering handle 51. The steering input is detected, for example, as a steering wheel angle or a hand torque, by means of a sensor 5 of the steering wheel module 2. The feedback torque 20 is generated by an actuator 6 of the steering wheel module 2 and applied to the steering handle 51.

[0068] The steering module 3 is configured to set a steering angle at at least one steerable wheel 52 based on the detected steering input and to determine the feedback torque 20. The required rack position of a rack 53 (or alternatively a connecting rod) is detected directly or indirectly by a sensor 7 of the steering module 3. The steering input is applied to the rack 53 and thereby to the wheels 52 by means of an electric motor 8 of the steering module 3. This is achieved in particular via a steering gear 11.

[0069] The steering input and the feedback torque 20 are each transmitted via the communication link 4.

[0070] Furthermore, the steering system 1 comprises a control unit 9, which in particular performs a control of the steering angle and the feedback torque 20 in a manner known per se. In particular, the control unit 9 is configured to control the steering angle and, for this purpose, to determine a manipulated variable, in particular a (target) torque for the electric machine 8. However, the control unit 9 can also be a separate control unit 9, which is configured solely for carrying out the method described in this disclosure.

[0071] The one in Fig. The embodiment shown in Figure 1 provides that the control unit 9 is arranged in the steering module 3. However, the control unit 9 can also be arranged in the steering wheel module 2 or outside of the steering wheel module 2 and the steering module 3.

[0072] The control unit 9 is configured to determine a non-linear kinematic tire trail value n R, in particular starting from at least one lateral acceleration detected and / or estimated at a steering axis of the steering system 1, furthermore based on the determined nonlinear kinematic tire trail i) to initiate feedback 21 at a steering handle 51 of the steering system 1 and / or ii) to control an assistance system function 56.

[0073] To determine the nonlinear kinematic tire trail value n R In a simple case, the control unit 9 determines, in particular starting from a detected or estimated lateral acceleration, a tire lateral force, and, starting from the tire lateral force, in particular a slip angle (in particular by means of a characteristic curve, cf. Fig. 7) and, starting from the slip angle determined in this way, the nonlinear kinematic tire caster value n R (especially by means of a characteristic curve, cf. Fig. 8) The procedure has already been described in more detail in the general description.

[0074] The feedback 21 is additionally (i.e., particularly as an additional moment) applied to the feedback moment 20. The assistance system function 56 is provided in particular by an assistance system 55 of the vehicle 50, for example, a lateral guidance assistant.

[0075] The Fig. Figure 2 shows a schematic flowchart to illustrate embodiments of the method and the steering. The control unit 9 is configured to detect and / or estimate a lateral acceleration. v on a steering axis of steering 1 and a detected engine torque M mot to obtain the electric machine 8. Starting from the measured or estimated lateral acceleration ÿ v The control unit 9 determines a tire side rudder F yv (Procedure step 100).

[0076] Furthermore, the control unit 9 determines, based on a non-linear kinematic tire trail value n R , a constructive tire trail value n K , a track lever ratio i kin and the specific tire lateral force F yv a first rack force F Zst,1 (Procedure step 101). The track lever ratio i kin For example, starting from a detected motor angle φ mot determined by placing it in a rack position x Zst (Procedure step 102) and by means of a track lever kinematics into the track lever ratio i kin (Procedure step 103) is converted.

[0077] Starting from the measured engine torque M mot The control device 9 also determines a second rack force F Zst,2 (Procedure step 104). The determined first rack force F Zst,1 and the specific second rack force F Zst,2are compared with each other (process step 105), and based on a comparison result, the determined tire lateral force F is calculated. yv using a correction factor c corr corrected to a corrected tire lateral force F yv,corr (Procedure step 106).

[0078] The control unit 9 is further configured to determine the nonlinear kinematic tire trail value n R starting from the corrected tire lateral force F yv,corr to determine (procedure steps 107 and 108). For this purpose, in particular an inverse tire model (e.g., using an empirically and / or by simulation determined characteristic curve of the slip angle α) is used. v above the tire sidewall F yv ) and a trailing model (e.g., using an empirically and / or by simulation determined characteristic curve of the nonlinear kinematic tire trailing value n) R above the slip angle α v) used, as already described in the general description.

[0079] The process steps are repeated, in particular, until the first rack force F is reached. Zst,1 has converged.

[0080] The specific first rack force F Zst,1 is provided by control unit 9.

[0081] It may be provided that a motor angular velocity φ̇ mot the electric machine 8 is detected, starting from the detected motor angular velocity φ̇ mot a friction compensation based on the motor torque M mot estimated second rack force F Zst,2 is carried out. For this purpose, starting from the measured motor angular velocity φ̇. mot a friction F Reib determined (process step 109) and the determined friction F Reib from the second rack force F Zst,2 subtracted (procedure step 110).

[0082] It may be provided that the lateral acceleration ÿ v by means of a lateral acceleration sensor 13 arranged on or near the steering axis ( Fig. 1) is recorded.

[0083] The lateral acceleration sensor 13 may be arranged in or on the steering module 3 of the steering system 1. In particular, the lateral acceleration sensor 13 may be arranged on a circuit board (PCB) of the control unit 9.

[0084] Alternatively or additionally, it may be provided that the lateral acceleration ÿ v The steering axis is estimated based on measured values ​​acquired by an inertial sensor 54 of the vehicle 50. A conversion is then carried out, in particular according to the procedure described in the general description.

[0085] It may be provided that a wheel contact force is detected and / or estimated, whereby a wheel contact mass m is calculated based on the detected and / or estimated wheel contact force. v is determined (procedure step 111), whereby the determined wheel contact mass m v when determining the tire lateral force F yv This is taken into account (procedure step 100).

[0086] It may also be provided that the wheel contact force is calculated based on recorded and / or estimated suspension travel and / or accelerations ẍ Fzg will be adjusted.

[0087] It may be provided that, starting from the first provided rack force F Zst,1 a feedback moment 20 ( Fig. 1) is intended and provided for the steering handle 51 of the steering system 1. This is done in a manner known per se.

[0088] It may be provided for, as in the Fig. Figure 2 schematically shows that the nonlinear kinematic tire trail value n R taking into account a deviation between a measured and / or estimated lateral acceleration ÿ v specific size and a starting from a measured motor torque M mot The specific quantity of the electric machine 8 used to apply a steering torque is determined. In particular, the determination and consideration are carried out in the process steps 105 and 106 shown.

[0089] It may also be provided for, in particular, as is done in the Fig. Figure 2 schematically shows that to determine the nonlinear kinematic tire trail value: - based on the measured and / or estimated lateral acceleration ÿ v a tire lateral force F yv is determined - starting from the non-linear kinematic tire trail value n R, a constructive tire trail value n K , a track lever ratio i kin and the specific tire lateral force F yv a first rack force F Zst,1 is determined - starting from a measured engine torque M mot The electric machine 8 used to apply a steering torque has a second rack force F Zst,2 is determined - the specific first rack force F Zst,1 and the specific second rack force F Zst,2 to be compared with each other, and - based on a comparison result, the specific tire lateral force F yv is corrected, whereby the nonlinear kinematic tire trail value n R starting from the corrected tire lateral force F yv,corr is determined.

[0090] The Fig. Figure 3 shows a schematic representation to clarify the quantities mentioned in this revelation and the relationships described in the general description. Here, δ v The steering angle of wheel 52 and x and y are Cartesian coordinates in a coordinate system of the vehicle.

[0091] The Fig. Figure 4 shows a schematic representation to clarify the quantities mentioned in this disclosure and the relationships described in the general description. In particular, a wheel contact area A and the nonlinear kinematic tire trail n are shown. R shown.

[0092] The Fig. Figure 5 shows a schematic representation to clarify the quantities mentioned in this disclosure and the embodiment in which the lateral acceleration is estimated from measured values ​​obtained from an inertial sensor. Here, l denotes hthe distance from the vehicle's center of gravity to the rear axle and l v the distance of the center of gravity to the front axle. The center of gravity position is given by S = (S x |S y |S z ) defined and the distance of the required virtual point of acceleration to the center of mass is given by P = (P x |P y |P z ) defined (so it is used in the representation in Fig. (5, assuming that S is located at the origin). Point P is shown here by way of example in the rear of the vehicle. In the method described in this disclosure, however, point P must be located in the steering system in order to calculate the lateral acceleration there.

[0093] The Fig. Figure 6 shows an example of a characteristic curve according to a Pacejka model, in which a tire lateral force F yv over a slip angle α v is shown.

[0094] The Fig. Figure 7 shows a characteristic curve that is inversely proportional to the one in the Fig. The characteristic curve shown in Figure 6 is (inverted Pacejka modeling). To determine the slip angle α v from the tire lateral force F yv As already described in the general description, only the dashed part up to the turning point is used, so that a slip angle can be determined for each tire lateral force.

[0095] The Fig. Figure 8 shows an example characteristic curve of the nonlinear kinematic tire trail value n R above the slip angle α v according to a Pacejka model. Such a characteristic curve is generated in process step 108 ( Fig. 2, Fig. 9 and Fig. 10) used to start from the slip angle α v the nonlinear kinematic tire trail value n R to determine, as already described in the general description. In the Fig. 8 is also the maximum value n R,max of the nonlinear kinematic tire trail value, which, within the framework of normalization, is used to form a ratio between this maximum value and the currently determined nonlinear kinematic tire trail value n. R is needed.

[0096] It may be provided that the strength of the feedback 21 and / or the control depends on a value of the specific nonlinear kinematic tire trail value n R is determined.

[0097] It may be provided that the specific non-linear tire caster value n R The normalized value is taken into account. This is done in particular by normalizing to the aforementioned maximum value n. R,max ( Fig. 8).

[0098] It may be provided that the dependence on the specific nonlinear kinematic tire trail value n Rtaking into account a given characteristic curve f(n) R ) is determined. In this context, the normalized value can be taken into account in particular.

[0099] It may be provided that the feedback and / or control only occurs after reaching a specific nonlinear kinematic tire trail value n. R The process occurs at a predetermined threshold. In particular, a threshold is specified that must be undercut in order for the (additional) feedback to be generated at the steering hand and / or for the assistance system function to be activated.

[0100] The Fig. Figure 9 shows a schematic representation of one embodiment of the method. Method steps 100 to 111 are the same as in the Fig. 2. In this embodiment, the (additional) feedback 21 is provided in the form of a periodic signal and / or an oscillation: FGrenz(t,FZst,1)=(1−cGrenz)⋅FA(FZst,1)⋅sin(2⋅π⋅t⋅f(FZst,1))

[0101] In a procedural step 112, the ratio c is determined in particular. Grenz formed, as already explained in the preceding description. In process step 113, starting from the ratio c Grenz the periodic signal is generated, in particular in the form of a periodically changing force F Grenz , which in process step 114 leads to the first rack force F Zst,1 is added, with the sum representing the total feedback force (see Eq. 14): FZst_Feedback=FZst,1+FGrenz

[0102] From this, the feedback torque 20 is then derived (not shown). Fig. 1) calculated at the steering handle and set or regulated there.

[0103] It may be provided that in process step 112 the first rack force F is also taken into account. Zst,1is taken into account, in particular to determine the amplitude of the periodic signal and / or the oscillation as a function of this first rack force F. Zst,1 to adjust.

[0104] The Fig. Figure 10 shows a schematic representation of an embodiment of the method. Method steps 100 to 111 are the same as in the Fig. 2 designed, process step 112 is as in the Fig. The embodiment shown in 9 is designed accordingly. In this embodiment, the feedback is provided in the form of a factor F. Gain The following is taken into account: a steering-induced feedback, represented here by the first rack force F. Zst,1 , scaled on the steering handle.

[0105] For this purpose, in a process step 113, starting from the ratio c determined in process step 113, Grenz the factor F Gain determined. This is done in particular with the help of a characteristic curve f.Gain (c Grenz The determined first rack force F Zst,1 is then used in process step 114 with the factor F Gain multiplied to determine the total feedback force: FZst_Feedback=FZst,1⋅fGain(cGrenz)

[0106] From this, the feedback torque 20 is then derived (not shown). Fig. 1) calculated at the steering handle and set or regulated there. Reference symbol list 1 Steering 2 Steering wheel module 3 Steering module 4 Communication link 5 Sensor 6 Actuator 7 Sensor 8 electric machine 9 Control unit 11 Steering gear 13 Lateral acceleration sensor 20 Feedback torque 21 Feedback (for vehicle dynamics limits) 50 vehicles 51 Steering handle 52 steerable wheel 53 Rack and pinion 54 Inertial sensors 55 Assistance system 56 Assistance system function 101-114 Procedural steps A wheel contact area c corr Correction factor c Grenz Ratio (normalized nonlinear kinematic tire tracking value) f(·) characteristic curve F Gain Factor (feedback) F Reib friction F yv Tire lateral force F yv,corr corrected tire lateral force F Zst,1 first rack force F Zst,2 second rack force F Zst_Feedback feedback i kin Track lever ratio M mot Motor torque m v Wheel contact mass n K Constructive tire trail value n R nonlinear kinematic tire trailing x Cartesian coordinate in the vehicle's coordinate system y Cartesian coordinate in the vehicle's coordinate system ẍ Fzg acceleration ÿ v Lateral acceleration (front axle) α v Slip angle φ mot Engine angle φ̇ mot Motor angle speed

Claims

[1] Method for determining and communicating a driving dynamics limit range of a steering system (1), where a nonlinear kinematic tire trail is instructed (n R ) is determined, where based on the determined nonlinear kinematic tire trailing (n R ) i) feedback (21) is generated at a steering handle (51) of the steering (1) and / or ii) an assistance system function (55) is activated; where to determine the nonlinear kinematic tire trail value (n R ): - based on the measured and / or estimated lateral acceleration (ÿ v ) a tire lateral force (F yv ) is determined, - starting from the nonlinear kinematic tire trail value (n R ), a constructive tire trail value (n K ), a track lever ratio (i kin ) and the specific tire lateral force (F yv) a first rack force (F Zst,1 ) is determined, - starting from a measured engine torque (M mot ) an electric machine (8) used to apply a steering torque a second rack force (F Zst,2 ) is determined, - the specific first rack force (F Zst,1 ) and the specific second rack force (F Zst,2 ) are compared with each other, and - based on a comparison result, the specific tire lateral force (F) yv ) is corrected, whereby the nonlinear kinematic tire trail value (n R ) starting from the corrected tire lateral force (F yv,corr ) is determined. [2] Method according to claim 1, characterized by , that the nonlinear kinematic tire trail value (n R ) taking into account a deviation between a lateral acceleration measured and / or estimated from the measured and / or estimated lateral acceleration (ÿ v) a certain size and a starting point based on a measured motor torque (M mot ) a certain size of an electric machine (8) used to apply a steering torque. [3] Method according to any of the preceding claims, characterized by , that a strength of the feedback (21) and / or the control depending on a value of the determined nonlinear kinematic tire trail value (n R ) is determined. [4] Method according to claim 3, characterized by , that the specific nonlinear kinematic tire trail is (n R ) is taken into account in a standardized way. [5] Method according to claim 3 or 4, characterized by , that the dependence on the specific nonlinear kinematic tire trail is (n R ) is determined taking into account a given characteristic curve (f(·)). [6] Method according to any of the preceding claims, characterized by, that the feedback (21) and / or the control only occurs after reaching a certain nonlinear kinematic tire trailing value (n) R ) specified threshold value. [7] Method according to any of the preceding claims, characterized by , that the feedback (21) takes the form of a periodic signal and / or an oscillation. [8] Method according to any of the preceding claims, characterized by , that the feedback (21) is in the form of a factor (F Gain ) is taken into account, which scales a steering-induced feedback at the steering handle (51). [9] Steering (1), encompassing: a control device (9), wherein the control device (9) is configured to determine a non-linear kinematic tire trail value (n) R ) to determine, furthermore based on the specific nonlinear kinematic tire trailing (n R ) i) to initiate feedback (21) at a steering handle (51) of the steering (1) and / or ii) to control an assistance system function (56); wherein the control device for determining the nonlinear kinematic tire trail value (n R ): - based on the measured and / or estimated lateral acceleration (ÿ v ) a tire lateral force (F yv ) certainly, - starting from the nonlinear kinematic tire trail value (n R ), a constructive tire trail value (n K ), a track lever ratio (i kin ) and the specific tire lateral force (F yv ) a first rack force (F Zst,1 ) certainly, - starting from a measured engine torque (M mot ) an electric machine (8) used to apply a steering torque a second rack force (F Zst,2 ) certainly, - the specific first rack force (FZst,1 ) and the specific second rack force (F Zst,2 compares them, and - based on a comparison result, the specific tire lateral force (F) yv ) corrected, and the nonlinear kinematic tire trail value (n R ) starting from the corrected tire lateral force (F yv,corr ) certainly.

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