JP4505484B2 - Variable steering angle ratio steering device for vehicle - Google Patents

Description

  The present invention relates to a variable steering angle ratio steering device for a vehicle capable of changing a ratio of a steering angle of a steered wheel to a steering angle of a steered wheel, that is, a steering angle ratio.

  Recently, in a steering apparatus for automobiles, one of means for improving operability is provided with a mechanism having a variable steering angle ratio. According to such a steering angle ratio variable mechanism, it is possible to reduce the force required for steering at a low speed range or a large steering angle range, and to improve running stability at a high speed range or a small steering angle range.

  Such a rudder angle ratio variable mechanism usually has a small rudder angle ratio in a large rudder angle region, that is, the steered wheel has a relatively large steered angle with respect to a given rotation angle of the steered wheel. In the small steering angle region, the steering angle ratio is large, that is, the steering angle of the steered wheels is relatively small with respect to the given rotation angle of the steered wheels.

In this way, by controlling the turning angle accurately in the high speed range where the turning angle is usually small, and enabling quick turning in the low speed range where the turning angle can be large, the vehicle can be used at any vehicle speed. A variable steering angle ratio steering device for a vehicle that can improve the handleability of the vehicle has already been proposed (for example, Patent Document 1).
JP 7-257406 A

  According to the knowledge of the inventor of the present application, in the above-described variable steering angle ratio steering device for a vehicle, when the steering angle ratio is further changed under a specific condition, the handleability of the vehicle can be further improved. For example, the steering angle ratio may be increased in order to suppress excessive turning of the steered wheels, particularly on road surfaces having a small friction coefficient, such as snowy surfaces, frozen road surfaces, and wet road surfaces.

  In view of the knowledge of the present inventor, the main object of the present invention is to provide a variable steering angle ratio steering device that further improves the handling of the vehicle, particularly in situations where it is desired to suppress excessive turning of the steering wheel. An object of the present invention is to provide a variable steering angle ratio steering device for a vehicle that can reduce the burden on the driver required for the control.

In order to achieve such an object, the variable steering angle ratio steering device for a vehicle according to the present invention can change a steering angle ratio that is a ratio of a steering input angle to a steered wheel and a steering output angle to a steerable wheel. A variable steering angle ratio steering device for a vehicle, wherein the input shaft is rotatably supported by a casing for inputting a steering input, and is rotatable to the casing for transmitting a steering output to steerable wheels. A supported output shaft, a steering angle ratio changing means provided between the input shaft and the output shaft to transmit a steering rotational force to the output shaft, and a maximum generated lateral force of the steerable wheel And the lateral force usage rate calculating means for calculating the lateral force usage rate, which is the ratio of the lateral force to the maximum generated lateral force, and the steering angle ratio to be achieved by the steering angle ratio changing means. and control means for determining, the control means, the side force utilization calculation Lateral force usage rate calculated by means increases the steering angle ratio when approached 1.

  In the vehicular variable rudder angle ratio steering apparatus according to the present invention, preferably, the lateral force usage rate detecting means detects a maximum generated lateral force based on a longitudinal force of a steerable wheel calculated based on a driving force and a braking force, and a steering force. The steerable wheel lateral force is calculated from the rack axial force of the apparatus and the steerable angle of the steerable wheel, and the ratio of the steerable wheel lateral force to the maximum generated lateral force is calculated as a lateral force usage rate.

  As described above, according to the present invention, the steering angle ratio is increased in accordance with the increase in the lateral force usage rate of the steered wheels to prevent overcutting, and the vehicle control burden on the driver can be reduced. .

  Embodiments of the present invention will be described in detail below based on specific examples shown in the accompanying drawings.

  FIG. 1 is a diagram showing an outline of a rack and pinion mechanism used in an automobile steering apparatus to which the present invention is applied. In FIG. 1, rotational motion (β) of a pinion 5 directly connected to a steering shaft 2 integrally coupled to a steering wheel (steering wheel) 1 via a connecting shaft 4 having a universal joint 3 is applied to the pinion 5. The linear motion of the rack 6 is converted into the linear motion (L) of the meshed rack 6, and the linear motion of the rack 6 is converted into the steering motion (θ) of the front wheels 9 that are steerable wheels via the tie rod 7 and the knuckle arm 8. It has become.

  2 to 4 show a portion indicated by V in FIG. 1 of the vehicle steering system to which the present invention is applied. In FIG. 2, the input connected to the steering wheel 1 (not shown). The shaft 11 is rotatably supported via a ball bearing 15 at an eccentric position of a support member 14 that is rotatably supported by the upper casing 13 a via a ball bearing 12. An expanding C-shaped coupling member 16 is formed at the lower end portion of the input shaft 11 that has entered the lower casing 13b, and a pair of inclined inner side surfaces 23 are defined on both sides thereof. Has been.

  In order to convert the rotational motion of the steering wheel 1 into a linear motion, the output shaft 17 integrally formed with the pinion 5 meshed with the rack 6 is rotatably supported by the lower casing 13b via a pair of ball bearings 18a and 18b. Has been. An intermediate shaft 19 protrudes at a position eccentric from the center of the upper end of the output shaft 17, is received in the coupling member 16 of the input shaft 11 that has entered the lower casing 13 b, and extends parallel to the output shaft 17. Exist. The intermediate shaft 19 and the coupling member 16 integrally formed with the input shaft 11 are connected to each other via a slider 21 with a flat needle bearing 20 interposed and a tapered roller bearing 22. The slider 21 has a box shape with an open bottom, and its side wall is expanded downward. The needle bearing 20 is disposed on the outer surfaces of the inclined side walls facing each other of the slider 21, and these side walls are parallel to the inclined inner side surface 23 of the coupling member 16.

  In this way, the lower surface of the coupling member 16 including the pair of flat needle bearings 20 is received in a groove defined by the inclined inner side surface 23 whose lower side is widened, with respect to the axis of the input shaft 11. It can slide freely in the orthogonal direction. An opening for receiving the intermediate shaft 19 is provided at the center of the upper wall of the slider 21 so as to be relatively rotatable via a tapered roller bearing 22.

  A seal member 36 having a flexible cylindrical portion is provided between the input shaft 11 and the upper casing 13a, thereby maintaining airtightness in the apparatus, and at the same time, the input shaft 11 is connected to the output shaft 17. A lateral movement with respect to the casing 13 due to the eccentricity is allowed.

  An adjusting screw 24 is screwed into an axial screw hole provided at the bottom of the lower casing 13 b, and the adjusting screw is in contact with an outer race of a ball bearing 18 b that supports the lower end of the output shaft 17. By appropriately tightening the adjustment screw 24, the pinion 5 is pressed in the axial direction, and an appropriate preload is applied between the input shaft 11 and the output shaft 17 via the coupling member 16. As a result, the backlash of the coupling assembly comprising the coupling member 16, the needle bearing 20 and the slider 21 can be removed to improve the connection rigidity.

  As shown well in FIG. 4, a fan-shaped partial worm wheel 25 is formed on a part of the outer peripheral portion of the support member 14. The worm wheel 25 is engaged with a worm 28 driven by a motor 27 via a worm speed reduction mechanism 26 so that the support member 14 can be rotated over a predetermined angular range. It has become.

  The worm 28 is supported by the upper casing 13a via a backlash removing member 29 using an eccentric cam, and a hexagonal bar wrench is engaged with a hexagon hole 30 provided at the end of the backlash removing member 29. By rotating them together with respect to the upper casing 13a, the shaft center moves so that the meshing gap with the worm wheel 25 changes.

  By removing the rubber or resin cap fitted in the opening provided in the upper casing 13a, the backlash removing member 29 can be exposed and the hexagonal hole 30 can be accessed. In addition, the worm 28 and the worm reduction mechanism 26 are connected via an Oldham coupling 31 in order to allow movement of the axis of the worm 28.

  A displacement sensor 33 made up of a differential transformer or the like is attached to the upper casing 13a in order to engage with a pin 32 projecting from the upper surface of the support member 14 and detect the rotation angle of the support member 14. . A vehicle speed sensor 34 for detecting the vehicle speed V is attached at an appropriate position of the vehicle.

  This apparatus inputs to the control device 35 the amount of rotation of the support member 14 generated by the displacement sensor 33, that is, the eccentricity signal for the output shaft 17 of the input shaft 11 supported by the displacement sensor 33 and the travel speed signal generated by the vehicle speed sensor 34. Then, the motor 27 is controlled by feedback control so that the target eccentricity set corresponding to the traveling speed matches the actual eccentricity from the displacement sensor 33. In addition, an output signal is input to the control device 35 from a steering angle sensor 36 that detects the steering angle α.

  The operation principle of the variable gear ratio steering apparatus configured as described above will be described below. As shown in FIGS. 2 to 4, since the support member 14 is eccentric with respect to the input shaft 11, a range in which the axis of the input shaft 11 is indicated by A <b> 0 to A <b> 2 as the support member 14 rotates. Moving. Thus, by appropriately changing the amount of eccentricity a between the input shaft 11 and the output shaft 17, a certain difference is produced between the rotation angles of the input shaft 11 and the output shaft 17, as will be described below. Can be made.

  By changing the amount of eccentricity a between the axis of the input shaft 11 and the axis of the output shaft 17 continuously over a range indicated by a0 to a2 (0 = a0 <a1 <a2), the output The ratio (α / β) of the rotation angle (input angle = α) of the input shaft 11 with respect to the rotation angle (output angle = β) of the shaft 17, that is, the steering angle ratio can be continuously changed.

  Next, the relationship between the input angle α and the output angle β will be described with reference to FIG. Here, the rotation center of the input shaft 11 is A, the rotation center of the output shaft 17 is B, the action point between the input shaft 11 and the intermediate shaft 19 is C, the dimension between BC is b, and the input shaft 11 and the output shaft. The displacement amount (size between AB) with respect to 17 is a, the rotation angle (steering wheel steering angle) of the input shaft 11 is α, and the rotation angle (pinion rotation angle) of the output shaft 17 is β. Let P be the position of the operating point between the input shaft 11 and the output shaft 17 when the input shaft 11 is rotated by an angle α.

  AP · sin α = b · sin β

  AP · cos α + a = b · cos β

  Because

  tan α = b · sin β / (b · cos β-a)

  That is,

α = tan ⁻¹ (b · sin β / b · cos β-a)
Is obtained.

  When the input shaft 11 is rotated by an angle α around the point A, the intermediate shaft 19 is rotated around the axis of the output shaft 17, that is, around the point B by the crank engagement with the slider 21 of the coupling member 16 of the input shaft 11. Turn angle β. As can be seen from FIG. 5, when the input angle α increases to α1 and the output angle β increases to β1, the point of action moves to Q, as will be explained in more detail below, for a given input angle increment. The output angle increment increases. Here, when the support member 14 is rotated and the shaft center of the output shaft 17 is at the point A0 and the points A and B coincide, that is, the input shaft 11 and the output shaft 17 are aligned with each other, They rotate at the same angle, that is, integrally. This relationship is shown by a one-dot chain line a0 in FIG.

  When the support member 14 is rotated and the axis of the input shaft 11 reaches the intermediate position, and the points A and B are in the general positions shown in FIG. 5, the output angle β, together with the input angle α, It changes as shown by the thick solid line a1 in FIG. That is, the change of the output angle β with respect to the given increment of the input angle α is small in the low speed region and increases as the input angle α increases. When the vehicle speed increases and the support member 14 is further rotated, the amount of eccentricity between the input shaft 11 and the output shaft 17 is maximized (the distance between the points A and B is maximized), and the thin solid line a2 in FIG. As shown by the above, such a tendency becomes more remarkable.

  In this way, in the low speed range, the rack stroke L with respect to the steering angle (input angle) α of the steered wheel is set to be larger than that of a conventional steering device, and more sensitive (sharp) characteristics can be realized. In the high speed range, the rack stroke L with respect to the steering angle α of the steered wheel can be made smaller than that of a conventional steering device, and more insensitive (stable) characteristics can be realized. Aside from the extremely low speed range, at a given vehicle speed, when the input angle is small, the output angle increment relative to the input angle increment is small, and as the input angle increases, the output angle increases relative to the input angle increment. The increment increases gradually. The relationship between the input steering angle and the rack stroke can be changed over the range indicated by the thick solid curve in FIG. 7 according to the vehicle speed.

  In short, according to the present invention, the relationship between the generally selected maximum steering angle (practical steering angle) and vehicle speed is expressed by a thin solid line curve (y) as shown by a thick solid line curve (y) in FIG. Compared to the corresponding relationship in the fixed steering angle ratio steering device indicated by x), it can be flattened.

  According to the first embodiment of the present invention, the controller 35 is responsive to the road friction coefficient μ. Such additional operating aspect features of the control device 35 are described below with respect to FIG.

  The control device 35 is usually composed of a microprocessor that operates according to a predetermined program, and includes functional means such as a road surface state estimation means 41, a steering angle ratio calculation means 42, or a steering angle ratio execution means 43. The road surface state estimation means 41 estimates the road surface friction coefficient μ by any known method, supplies the estimated road surface friction coefficient μ to the steering angle ratio calculation means 42, and the steering angle ratio calculation means 42 includes a vehicle speed sensor. The standard steering angle ratio command value K based on the vehicle speed V from the vehicle and the steering angle α which is the steering steering angle, and the standard steering angle ratio command value K based on the estimated road friction coefficient μ are corrected. The correction coefficient R for changing to is calculated. The steering angle ratio execution means 43 drives the motor 27 to realize a corrected steering angle ratio command value K *.

  Usually, when the road surface friction coefficient μ is relatively small, it is desirable to increase the steering angle ratio in order to avoid excessive steering. As shown in FIG. 10, the rudder angle ratio calculating means 42 has a correction coefficient R to be multiplied by the reference rudder angle ratio command value K to obtain the corrected rudder angle ratio command value K *, and the road surface friction coefficient μ. A correction coefficient map M1 given as a function is incorporated. As shown in FIG. 10, the correction coefficient R has a minimum value of 1 and increases as the road friction coefficient μ decreases.

  For the estimation method of the road surface friction coefficient μ, refer to Japanese Patent Application Laid-Open Nos. 9-058514 and 9-281030 of the same applicant. Here, an example of how to obtain the road surface state estimated value μ is shown below.

  The road surface state estimated value μ is related to the cornering power Cp of the tire from the FIALA equation (up to the second term) and expressed as follows.

  Cp = Sc (1−0.0166Sc / μW)

  However, Sc: Cornering stiffness

  μ: Estimated road surface condition

  W: Contact load

  Therefore, as shown in FIG. 11, the tire cornering power Cp decreases as the road surface state estimated value μ decreases. In the case of the rack / pinion type steering device, the rack axial force Fr at the same rudder angle is determined by the road surface state. It becomes smaller as the estimated value μ decreases. Therefore, the estimated road surface condition value μ is an actual rack shaft reaction force Frc with respect to the front wheel steering angle δ, and a reference rack shaft reaction force Frm set in advance as an internal model based on the vehicle design value and the result of the experimental measurement. Can be estimated.

  That is, since the cornering power Cp of the tire decreases as the road surface state estimation value μ decreases (see FIG. 11), in the case of a rack / pinion type steering device, the rack axial reaction force received from the road surface at the same steering angle is It becomes smaller as the estimated state value μ decreases. Therefore, if the front wheel rudder angle and rack shaft reaction force are measured and the actual rack shaft reaction force with respect to the front wheel rudder angle is compared with the reference rack shaft reaction force set in advance as an internal model, the estimated road surface condition value μ is estimated. can do.

  When the electric steering force assisting device is used, the friction coefficient between the road surface and the tire can be estimated based on, for example, the output signal of the axial force sensor incorporated in the device.

  Next, an internal model serving as a comparison reference for the actual rack shaft reaction force Frc is set as follows. As shown in FIG. 12, the steering angle α input from the steered wheels is converted into a rack shaft stroke amount via a transmission ratio Np with the pinion. A front wheel side slip angle γ is generated according to the stroke amount of the rack shaft. Here, the transfer function Gβ (s) of the front wheel side slip angle γ with respect to the stroke amount of the rack shaft changes due to the change of the stability factor accompanying the change of the road surface state estimated value μ.

  By applying the cornering power Cp and the trail ζ (caster trail + pneumatic trail) to the side slip angle γ of the front wheel, a moment around the kingpin is obtained. Here, the cornering power Cp and the pneumatic trail vary depending on the road surface state estimated value μ and the ground load W. A model rack shaft reaction force Frm is obtained by dividing the moment around the kingpin by the distance between the tire rotation center and the rack shaft center, that is, the knuckle arm length rk.

  From the above, the response of the model rack shaft reaction force Frm to the steering wheel steering angle α can be described by one transfer function Gf (s) derived from the calculation result based on each specification or the identification result from the actual vehicle measurement value. I understand that there is.

  From the actual rack shaft reaction force value Frc and the model rack shaft reaction force value Frm obtained as described above, the actual and model rack shaft reaction force increases with respect to the increase Δα of the steering wheel steering angle α are obtained (see FIG. 13). In a rudder angle range in which the vehicle response is linearly approximated, a ratio ΔFrc / ΔFrm between the actual rack shaft reaction force increase rate ΔFrc / Δα and the model rack shaft reaction force increase rate ΔFrm / Δα is set in advance. The road surface state estimated value μ can be estimated with reference to the road surface state estimated value determination map (FIG. 14).

  The rudder angle ratio calculating means 42 calculates a reference rudder angle ratio command value K that can perform stable cornering based on the vehicle speed V and the steering angle α, and based on the road surface state estimated value μ, the rudder angle ratio map M1. A correction coefficient R as a correction value of the reference rudder angle ratio according to the road surface condition is obtained, and both are multiplied to calculate a corrected rudder angle ratio command value K *. Therefore, steering can be performed with a steering angle ratio (gear ratio) in consideration of the road surface μ, and it is possible to suppress excessive steering in a low road surface μ state such as running on snow.

  Next, as a second embodiment of the present invention, control for changing the steering angle ratio in accordance with the front wheel lateral force usage rate (margin) will be described below with reference to the block diagram of FIG. In FIG. 15, the same parts as those in the illustrated example are denoted by the same reference numerals, and detailed description thereof is omitted.

  In this control, the road surface state estimated value μ from the road surface state estimating means 41 similar to the above is input to the equivalent friction circle setting means 45, and the equivalent friction circle data from the equivalent friction circle setting means 45 is converted into the lateral force. Input to the usage rate calculation means 46. Regarding the setting of the equivalent friction circle, the equivalent friction circle is set based on the estimated road surface state value μ, and the shape of the equivalent friction circle (ellipse) is selected from data of a plurality of equivalent friction circles stored in advance.

  The lateral force usage rate calculating means 46 includes, for example, a driving force obtained from the relationship between the vehicle speed and the intake pipe negative pressure, a front wheel longitudinal force Fx calculated based on a braking force obtained from the brake fluid pressure, and a rack (not shown). The front wheel lateral force (cornering force) Fy calculated based on the rack axial force by the axial force detection means and the front wheel steering angle is input.

  The lateral force usage rate calculating means 46, which is a lateral force usage rate detecting means, calculates the front wheel lateral force usage rate ξ based on the input data and outputs it to the steering angle ratio calculating means 47. The front wheel lateral force usage rate ξ can be calculated as follows. That is, the maximum frictional force generated by the tire can be calculated from the set equivalent friction circle, and the maximum generated lateral force can be obtained by subtracting the amount of frictional force used by the longitudinal force. The ratio of the lateral force to the maximum generated lateral force is set as the lateral force usage rate.

  Similarly to the above, the vehicle speed V and the steering angle α are input to the steering angle ratio calculation means 47, the steering angle ratio command value K is calculated based on the vehicle speed V and the steering angle α, and the front wheel lateral force usage rate ξ Based on the above, the correction coefficient R for converting the reference steering angle ratio command value K into the corrected steering angle ratio command value K * is calculated, and the obtained corrected steering angle ratio command value K * is output to the steering angle ratio execution means 43. The front wheel 9 is steered based on the vehicle speed and in consideration of the front wheel lateral force usage rate ξ. As shown in FIG. 16, in the steering angle ratio calculation means 47, as the front wheel lateral force utilization rate ξ approaches 1, that is, as the front wheel lateral force margin decreases, the correction coefficient R increases greatly. A steering angle ratio map M2 set as described above is incorporated. Thus, as the front wheel lateral force usage rate ξ approaches 1, that is, as the front wheel lateral force margin becomes smaller, the front wheel lateral force usage rate ξ exceeds 1 by increasing the steering angle ratio. Can be avoided. The control can be similarly performed by using the lateral G usage rate calculated based on the lateral G detected by providing the lateral G sensor instead of the front wheel lateral force usage rate ξ.

  According to the third embodiment of the present invention, the control device responds to the position of the vehicle with respect to the travel lane. In order to perform this control, a CCD camera is attached in the vicinity of the rear-view mirror in the upper part of the driver's seat of the vehicle. A traveling lane is extracted, and the position and direction of the own vehicle in the lane are determined from the image data. This control procedure will be described below with reference to the vehicle travel explanatory diagram of FIG. 17 and the block diagram of FIG.

  First, the road shape detection unit 51 performs normal image processing on the road shape in the vehicle traveling direction taken by the CCD camera, and extracts a lane RL as a road shape detection target in this control. Information on the road shape (lane RL) is input to the maximum lateral movement amount calculation means 52. Instead of the CCD camera, a GPS or other navigation system provided with map information can be used.

  The maximum lateral movement amount calculation means 52 obtains a lateral movement range that is possible for the vehicle W to stay in both the left and right lanes RL. Note that the lateral movement range covers both the left and right sides of the vehicle W, but the right side in FIG. The maximum lateral movement amount calculation means 52 calculates the distance between the right end of the vehicle W and the lane RL on the right side thereof from the detected value of the road shape, and uses this as the maximum lateral movement amount Smax to calculate the steering angle ratio. Output to means 53.

  Further, the predicted lateral movement amount calculation means 54 obtains the predicted lateral movement amount Sexp of the vehicle W (imaginary line in FIG. 17) after t seconds and outputs it to the steering angle ratio calculation means 53. The lateral movement amount predicted value Sexp can be calculated from the detected values of the vehicle speed V and the turning angle θ. The time t is set according to the vehicle speed and the vehicle characteristics, but may be set to, for example, around 1 second.

  The steering angle ratio calculation means 53 calculates a reference steering angle ratio command value K based on the vehicle speed V and the turning angle θ, and multiplies it by a correction coefficient R. The parameter for selecting the correction coefficient R is based on the ratio of the maximum lateral movement amount Smax and the predicted lateral movement amount Sexp. In this case, as shown in FIG. 19, the map M3 therefor is set so that the correction coefficient R is gradually increased as the ratio approaches 1.

  Thus, by increasing the steering angle ratio as the ratio of the expected lateral displacement of the vehicle to the distance between the lane marking and the vehicle increases, the vehicle is prevented from deviating from the travel lane. The This gives the vehicle a tendency to stay in the travel lane, which means that the driver can maintain the travel lane with less effort compared to conventional vehicles.

  The fourth embodiment shown in FIGS. 20 and 21 can achieve the same object with a slightly different configuration. First, the front of the vehicle is monitored by a CCD camera or the like, and the target lateral position Ptar of the vehicle after t seconds (which may be a lateral distance with respect to the lane RL) is obtained. This is compared with the predicted lateral position value Pexp of the vehicle after t seconds calculated from the detected values of the vehicle speed V and the turning angle θ. The steering angle ratio calculation means 53 calculates a reference steering angle ratio command value K based on the vehicle speed V and the turning angle θ, and multiplies it by a correction coefficient R. The parameter for selecting the correction coefficient R is based on the deviation between the target lateral position Ptar and the predicted lateral position value Pexp. In this case, as shown in FIG. 22, the map is set so that the correction coefficient R is gradually increased as the deviation becomes smaller. Also in this case, when the vehicle is expected to deviate from the travel lane, the vehicle is prevented from deviating from the travel lane by increasing the steering angle ratio. This gives the vehicle a tendency to stay in the travel lane, which means that the driver can maintain the travel lane with less effort compared to conventional vehicles.

1 is a schematic configuration diagram of a vehicle steering apparatus to which the present invention is applied. Sectional drawing which shows the structure of the variable gear ratio steering apparatus with which this invention was applied. The exploded perspective view of a shaft part. Sectional drawing which follows the IV-IV line | wire of FIG. Explanatory drawing for demonstrating the principle of operation of this variable gear ratio steering apparatus. The steering angle ratio characteristic diagram of this variable gear ratio steering apparatus. Steering angle ratio characteristic diagram. The graph which shows the relationship between driving speed and steering wheel rotation angle. The block diagram of the control according to the state of road surface in the 1st example of the present invention. The figure which shows the control point in the rudder angle ratio calculation means. The relationship diagram of cornering power and a road surface friction coefficient. Flow chart related to the setting of the internal model. The increase line figure of the vehicle state quantity with respect to the amount of steering angles. Judgment map for road friction coefficient. The block diagram of the control which changes a steering angle ratio according to the front wheel lateral force usage rate in 2nd Example of this invention. The figure which shows the control point in the steering angle ratio calculation means 47. FIG. The vehicle travel explanatory drawing in 3rd Example of this invention. The block diagram for performing lane departure suppression control. The figure which shows the control point in the steering angle ratio calculation means 53. FIG. The vehicle travel explanatory drawing in 4th Example of this invention. The block diagram for performing lane departure suppression control. The figure which shows the control point in the steering angle ratio calculation means 53. FIG.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Steering wheel 11 Input shaft 13a Upper casing 13b Lower casing 14 Support member 16 Coupling member 17 Output shaft 19 Intermediate shaft 35 Controller 41 Road surface state estimation means 42 Steering angle ratio calculation means 43 Steering angle ratio execution means 44 Vehicle 45 Equivalent friction circle Setting means 46 Lateral force usage rate calculating means 47 Steering angle ratio calculating means 51 Road shape detecting means 52 Maximum lateral movement amount calculating means 53 Steering angle ratio calculating means 54 Lateral movement amount predicted value calculating means

Claims (2)

A variable steering angle ratio steering device for a vehicle capable of changing a steering angle ratio, which is a ratio of a steering input angle to a steered wheel and a steering output angle with respect to a steerable wheel,
An input shaft rotatably supported by the casing for inputting steering input;
An output shaft rotatably supported by the casing for transmitting steering output to the steerable wheel;
Rudder angle ratio changing means provided between the input shaft and the output shaft in order to transmit a steering torque to the output shaft;
A lateral force usage rate calculating means for calculating the maximum generated lateral force and lateral force of the steerable wheel and calculating a lateral force usage rate which is a ratio of the lateral force to the maximum generated lateral force ;
Control means for determining a steering angle ratio to be achieved by the steering angle ratio changing means,
It said control means, steering apparatus characterized by increasing the steering angle ratio when the lateral force usage rate calculated by said lateral force usage rate calculation means is close to 1. The lateral force usage rate calculating means calculates the maximum frictional force generated by the tire by an equivalent friction circle set based on the estimated value of the road surface friction coefficient, and based on the driving force and the braking force, the longitudinal force of the steerable wheels is calculated. And the maximum generated lateral force is calculated by subtracting the longitudinal force from the calculated maximum frictional force, and the lateral force of the steerable wheel is calculated from the rack axial force of the steering device and the steering angle of the steerable wheel, The steering device according to claim 1, wherein a lateral force usage rate is calculated from the calculated maximum generated lateral force and lateral force .

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