JP6634877B2 - Steering control device - Google Patents

Description

  The present invention relates to a steering control device.

  2. Description of the Related Art Conventionally, there has been known an electric power steering device (hereinafter, referred to as “EPS”) that assists a driver in steering by applying power of a motor to a vehicle steering mechanism. As an example of the steering mechanism, there is a so-called rack and pinion type. The turning angle of the steered wheels is changed by converting the rotational motion of the steering shaft into a linear motion of the rack shaft through meshing between the pinion shaft and the rack shaft. The EPS control device controls the motor based on the detection results of various sensors provided in the vehicle.

  For example, an EPS control device described in Patent Literature 1 calculates a basic assist component that is a basic component of a steering assist force to be applied to a steering mechanism based on a steering torque and a vehicle speed. Further, the control device calculates a target pinion angle based on the basic drive torque which is a sum of the steering torque and the basic assist component, and calculates a correction component for the basic assist component through feedback control that matches the actual pinion angle with the target pinion angle. Calculate.

  The control device calculates a spring component of the basic drive torque based on the target pinion angle in order to increase the steering reaction force in accordance with the increase in the lateral acceleration acting on the vehicle, and calculates the magnitude of the basic drive torque in accordance with the spring component. After the reduction, a correction component for the basic assist component is calculated. The control device calculates a first component of the spring component based on the target pinion angle and a second component based on the lateral acceleration, and sets a use ratio of the first component and the second component according to the magnitude of the lateral acceleration. The control device decreases the usage ratio of the first component as the lateral acceleration increases, and increases the usage ratio of the second component.

  The magnitude of the target pinion angle, and consequently, the correction component for the basic assist component is reduced by the amount by which the basic drive torque is reduced according to the spring component. As the correction component decreases, the steering assist force applied to the steering mechanism also decreases. Since the steering reaction force increases as the steering assist force decreases, a suitable steering reaction force can be obtained according to the magnitude of the lateral acceleration. Therefore, it is possible to harmonize the lateral acceleration received by the driver with the body and the steering reaction force (response) felt by the hand through the steering.

  In addition, the first component of the spring component based on the target pinion angle contributes to a so-called rigid feeling (a firm feeling). The second component of the spring component based on the lateral acceleration contributes to a sense of unity with the vehicle. For this reason, by adjusting the use ratio of the first component and the second component according to the lateral acceleration, a sense of rigidity and a sense of unity can be suitably obtained. For example, it is conceivable that the greater the lateral acceleration, the less the sense of unity with the vehicle will be. Thereby, a sense of unity with the vehicle can be suitably obtained according to the magnitude of the lateral acceleration.

JP 2015-42528 A

  In the control device of Patent Literature 1, when the usage ratio of the first component of the spring component and the second component with respect to the first component is set according to the magnitude of the lateral acceleration, the following is a concern. . For example, when the vehicle is traveling on a low friction road, lateral acceleration is unlikely to occur. Since the use ratio of the second component decreases as the lateral acceleration decreases, the spring component of the basic drive torque tends to be mainly the first component when the vehicle is traveling on a low friction road. Since the first component of the spring component contributes to the so-called rigidity, the driver feels a stronger rigidity as a response. There is a possibility that the driver may feel uncomfortable with the fact that a firm response is obtained on a low friction road where it is difficult for the driver to feel a firm response as a steering feel.

  An object of the present invention is to provide a steering control device capable of more appropriately transmitting a road surface condition to a driver as a steering reaction force.

  A steering control device that solves the above-described problem is a steering control device that controls a motor that is a source of a driving force applied to a steering mechanism of a vehicle based on an assist command value calculated in accordance with a steering operation. A first calculating unit that calculates a basic component of the assist command value according to a steering torque; and a feedback that matches an actual rotation angle of a rotating body rotating through a steering operation with a target rotation angle calculated according to the basic component. A second calculation unit that calculates a compensation component for the basic component through control. The second calculation unit calculates at least a first component based on the target rotation angle and a second component based on a state quantity reflecting vehicle behavior or a road surface state, and calculates the first component and the It is assumed that the target rotation angle is calculated in accordance with the basic component corrected based on the value added to the second component. The second arithmetic unit is configured to, when a lateral acceleration difference value that is a difference between an estimated lateral acceleration based on a state quantity reflecting a vehicle behavior or a road surface state and an actual lateral acceleration detected through an on-vehicle sensor increases, There is an arithmetic function part for increasing the use ratio of the second component than before the acceleration difference value increases.

  As a situation where the lateral acceleration difference value is likely to increase, for example, a situation where the vehicle is traveling on a low friction road is assumed. Here, the first component based on the target rotation angle is a component that contributes to the sense of rigidity of the steering. The second component based on the state quantity reflecting the vehicle behavior is a component that contributes to a sense of unity with the vehicle. Therefore, for example, when the vehicle is traveling on a low friction road and the usage ratio of the first component is larger than the usage ratio of the second component, the sense of rigidity of the steering is superior to the sense of unity with the vehicle. . In this case, since the road surface reaction force is smaller on the low friction road, there is a possibility that the steering feeling may be perceived as a response despite the fact that it is originally difficult to feel a firm response as the steering feel. In this regard, according to the above configuration, when the lateral acceleration difference value increases, the use ratio of the second component is increased as compared with before the lateral acceleration difference value increases, and the second component is increased in accordance with the increase. The use ratio of the component 1 is reduced. For this reason, the driver can easily recognize that the vehicle is traveling on a low friction road by feeling that the sense of rigidity has decreased as a response.

In the steering control device described above, it is preferable that the second calculation unit increases the usage ratio of the second component as the lateral acceleration difference value increases.
According to this configuration, the usage ratio of the second component is more finely increased or decreased according to the vehicle behavior or the road surface condition. The driver feels a finer change in the sense of rigidity as a response, thereby making it easier to grasp the vehicle behavior or the road surface condition.

  In the above steering control device, the steering mechanism has a pinion shaft that rotates in conjunction with a steering operation, and a turning shaft that steers steered wheels in conjunction with the pinion shaft. In this case, a rack axial force calculated based on the value of the current supplied to the motor and the steering torque may be calculated as the second component.

  As the road surface reaction force increases, the rack axial force increases. Conversely, the smaller the road surface reaction force, the smaller the rack axial force. As described above, the rack axial force can be used as a state quantity reflecting the road surface reaction force.

  In the steering control device described above, the second calculation unit may include a first reaction force component based on the target rotation angle, and a second reaction force based on a yaw rate and a lateral acceleration, which are state quantities reflecting vehicle behavior. Assuming that the first component is calculated by adding the components at a usage ratio determined according to the actual lateral acceleration, the second calculating unit determines that the second counter unit increases as the actual lateral acceleration increases. It is preferable to increase the use ratio of the force component.

  According to this configuration, the first component is calculated by adding the first reaction force component and the second reaction force component. In the first component, the use ratio of the second reaction force component increases as the lateral acceleration increases, and the use ratio of the first reaction force component decreases in accordance with the increase. Here, the first reaction force component based on the target rotation angle is a component that contributes to the sense of rigidity of the steering. The second reaction force component based on the state quantity reflecting the road surface reaction force is a component that contributes to a sense of unity with the vehicle. For this reason, as the lateral acceleration increases, the sense of unity with the vehicle becomes superior to the sense of rigidity of the steering. Conversely, as the lateral acceleration decreases, the sense of rigidity of the steering becomes superior to the sense of unity with the vehicle. Therefore, when the lateral acceleration difference value is larger, the second component that contributes to the sense of unity with the vehicle is used in preference to the first component, but when the vehicle is traveling on a flat dry road, When the lateral acceleration difference value is smaller, the first component that contributes to the sense of rigidity of the steering is used in preference to the second component. Therefore, when the vehicle is gripping on a dry road, the driver feels the sense of rigidity as a steering reaction force more strongly. Since the driver can feel a sharp response, it is easy to drive.

  In the steering control device described above, the second calculation unit may include a first reaction force component based on the target rotation angle and a second reaction force component based on a yaw rate and a lateral acceleration, which are state quantities reflecting vehicle behavior. Is calculated based on the actual lateral acceleration or the usage ratio determined according to the lateral acceleration difference value. In addition, it is preferable that the use ratio of the second reaction force component is increased as the actual lateral acceleration or the lateral acceleration difference value, of which the increasing tendency is larger among the increasing trends of the lateral acceleration difference value, is increased.

  According to this configuration, the first component is calculated by adding the first reaction force component and the second reaction force component. For example, when the vehicle is traveling on a low friction road, lateral acceleration is less likely to occur, but the lateral acceleration difference value tends to be larger. For this reason, when the vehicle is traveling on a low friction road, the tendency of increasing the lateral acceleration difference value tends to be larger than the tendency of increasing the actual lateral acceleration. In this case, as the lateral acceleration difference value increases, the usage ratio of the second reaction force component increases, and the usage ratio of the first reaction force component decreases in accordance with the increase. Therefore, the first component is a component that contributes to the sense of unity with the vehicle as a whole. Here, as the lateral acceleration difference value increases, the use ratio of the second component increases, but the second component also contributes to the sense of unity with the vehicle. Therefore, when the lateral acceleration difference value increases while traveling on a low friction road or the like, a sense of unity with the vehicle according to the increase in the lateral acceleration difference value is obtained.

  In the above steering control device, the second calculation unit calculates a first reaction force component based on the target rotation angle as the first component, and reflects a vehicle behavior as the second component. A second reaction force component based on the yaw rate and the lateral acceleration, which are state quantities, is calculated, and the first reaction force component and the second reaction force component are determined according to the actual lateral acceleration or the lateral acceleration difference value. Assuming that a spring component as a correction value for the basic component is calculated by adding up the use ratios, the second calculating unit calculates the tendency of the actual lateral acceleration and the tendency of the lateral acceleration difference value to increase. It is preferable that the use ratio of the second reaction force component is increased as the actual lateral acceleration or the lateral acceleration difference value having a larger increasing tendency increases.

  According to this configuration, the spring component as a correction value for the basic component is calculated by adding the first reaction force component (first component) and the second reaction force component (first component). Is done. For example, when the vehicle is traveling on a low friction road, lateral acceleration is less likely to occur, but the lateral acceleration difference value tends to be larger. For this reason, when the vehicle is traveling on a low friction road, the tendency of increasing the lateral acceleration difference value tends to be larger than the tendency of increasing the actual lateral acceleration. In this case, as the lateral acceleration difference value increases, the usage ratio of the second reaction force component increases, and the usage ratio of the first reaction force component decreases in accordance with the increase. Therefore, the spring component as a correction value for the basic component is a component that contributes to the sense of unity with the vehicle as a whole. Further, the sense of rigidity of the steering is reduced by the reduction in the usage ratio of the first reaction force component. The driver can easily recognize that the vehicle is traveling on a low friction road by feeling that the sense of rigidity has decreased in response.

  ADVANTAGE OF THE INVENTION According to the steering control apparatus of this invention, a road surface state can be more appropriately transmitted to a driver as a steering reaction force.

FIG. 1 is a block diagram showing a configuration of an electric power steering device on which an electronic control device according to a first embodiment is mounted. FIG. 2 is a control block diagram of the electronic control device according to the first embodiment. FIG. 3 is a control block diagram of a target pinion angle calculation unit according to the first embodiment. FIG. 2 is a control block diagram of the ideal vehicle model according to the first embodiment. 5 is a graph showing a first gain map according to the first embodiment. 5 is a graph showing a second gain map according to the first embodiment. FIG. 10 is a control block diagram of an electronic control device according to a second embodiment. FIG. 10 is a control block diagram of an ideal vehicle model according to the second embodiment. 9 is a graph showing a third gain map according to the second embodiment. FIG. 13 is a control block diagram of an ideal vehicle model according to a third embodiment. FIG. 14 is a control block diagram of an ideal vehicle model according to a fourth embodiment. 14 is a graph showing a fourth gain map according to the fourth embodiment. FIG. 14 is a control block diagram of an electronic control device according to a fifth embodiment.

<First embodiment>
Hereinafter, a first embodiment in which the steering control device is embodied as an electric power steering device for a vehicle will be described.

<Overview of electric power steering device>
As shown in FIG. 1, an electric power steering device (EPS) 10 includes a steering mechanism 20 that turns a steered wheel based on a driver's steering operation, a steering assist mechanism 30 that assists the driver's steering operation, and An electronic control unit (ECU) 40 for controlling the operation of the steering assist mechanism 30 is provided.

The steering mechanism 20 includes a steering wheel 21 operated by a driver, and a steering shaft 22 that rotates integrally with the steering wheel 21. The steering shaft 22 includes a column shaft 22a connected to the steering wheel 21, an intermediate shaft 22b connected to a lower end of the column shaft 22a, and a pinion shaft 22c connected to a lower end of the intermediate shaft 22b. I have. The lower end of the pinion shaft 22c is engaged with a rack shaft 23 (more precisely, a portion 23a having rack teeth) extending in a direction intersecting the pinion shaft 22c. Therefore, the rotational movement of the steering shaft 22 is converted into a reciprocating linear movement of the rack shaft 23 by the rack and pinion mechanism 24 including the pinion shaft 22c and the rack shaft 23. The reciprocating linear motion, by being transferred respectively to the left and right steered wheels 26, 26 via tie rods 25 connected to both ends of the rack shaft 23, changes the steered angle theta _ta of steered wheels 26, 26 Is done. The traveling direction of the vehicle is changed by the turning angle theta _ta of the steered wheels 26 is changed.

  The steering assist mechanism 30 includes a motor 31 that is a source of a steering assist force. As the motor 31, a three-phase AC motor such as a brushless motor is used. The motor 31 is connected to the column shaft 22a via a speed reduction mechanism 32. The reduction mechanism 32 reduces the rotation of the motor 31 and transmits the reduced rotation force to the column shaft 22a. That is, by applying the motor torque to the steering shaft 22 as a steering assist force (assist force), the driver's steering operation is assisted.

The ECU 40 acquires detection results of various sensors provided in the vehicle as information indicating a driver's request or a traveling state, and controls the motor 31 according to the acquired various information. Various sensors include, for example, a vehicle speed sensor 410, a torque sensor 420, a rotation angle sensor 430, a lateral acceleration sensor 440, and a yaw rate sensor 450. The vehicle speed sensor 410 detects a vehicle speed (running speed of the vehicle) V. The torque sensor 420 is provided on the column shaft 22a, detects the steering torque T _h applied to the steering shaft 22 via the steering wheel 21. Rotation angle sensor 430 is provided in the motor 31 detects the rotation angle theta _m of the motor 31. Lateral acceleration sensor 440 detects lateral acceleration LA acting on the vehicle. The lateral acceleration LA refers to an acceleration acting in a left-right direction orthogonal to the traveling direction of the vehicle when the vehicle is viewed from above. Yaw rate sensor 450 detects yaw rate YR. The yaw rate YR refers to a rotational angular velocity around a vertical axis passing through the center of gravity of the vehicle (the rate of change of the rotational angle in the turning direction of the vehicle). ECU40 the vehicle speed V is acquired through these sensors, the steering torque _{T h,} the rotation angle theta _m, based on the lateral acceleration LA and yaw rate YR, and controls the motor 31.

<Schematic configuration of ECU>
Next, the hardware configuration of the ECU will be described.
As shown in FIG. 2, the ECU 40 includes an inverter circuit 41 and a microcomputer 42.

  The inverter circuit 41 converts a DC current supplied from a DC power supply such as a battery into a three-phase AC current based on a motor drive signal described later generated by the microcomputer 42. The converted three-phase AC current is supplied to the motor 31 via the power supply path 44 of each phase. A current sensor 45 is provided in the power supply path 44 of each phase. These current sensors 45 detect an actual current value I generated in the power supply path 44 of each phase. In FIG. 2, the power supply path 44 of each phase and the current sensor 45 of each phase are collectively illustrated as one for convenience of explanation.

The microcomputer 42 captures the detection results of the vehicle speed sensor 410, the torque sensor 420, the rotation angle sensor 430, the lateral acceleration sensor 440, the yaw rate sensor 450, and the current sensor 45 at a predetermined sampling cycle. Microcomputer 42, these detection results to be incorporated, i.e. the vehicle speed V, the steering torque _{T h,} to produce a rotation angle theta _m, the lateral acceleration LA, the motor drive signal based on the yaw rate YR and the current value I (PWM driving signal).

  To be precise, the microcomputer 42 performs the vector control of the motor current through the PWM drive of the inverter circuit 41. Vector control separates the motor current into a d-axis component (field current component) parallel to the magnetic field and a q-axis component (torque current component) orthogonal to the magnetic field. Is what you do. By the vector control, the motor 31 can be handled similarly to the DC motor.

<Microcomputer>
Next, a functional configuration of the microcomputer will be described.
The microcomputer 42 has various arithmetic processing units realized by executing a control program stored in a storage device (not shown). As shown in FIG. 2, the microcomputer 42 includes an assist command value calculation unit 51, a current command value calculation unit 52, a motor drive signal generation unit 53, and a pinion angle calculation unit 54 as these calculation processing units.

The assist command value calculation unit 51 captures the vehicle speed V, the steering torque _Th , and a pinion angle θ _p described later calculated by the pinion angle calculation unit 54, respectively, and based on these various types of captured information, the assist command value _Ta. ^* Is calculated. The assist command value T _a ^* is a command value indicating the rotational force to be generated in the motor 31 (assist torque).

The current command value calculator 52 calculates _a current command value I ^* based on the assist command value _Ta ^* calculated by the assist command value calculator 51. The current command value I ^* is a command value indicating a current to be supplied to the motor 31. To be precise, the current command value I ^* includes a q-axis current command value and a d-axis current command value in the d / q coordinate system. d / q coordinate system is a rotating coordinate according to the rotation angle theta _m of the motor 31.

Motor driving signal generator 53, a current command value I ^{* and} the actual current value I, and takes in the rotation angle theta _m of the motor 31, respectively, the actual current value I on the basis of these information captured in the current command value I ^* The current feedback control is performed so as to follow. The motor drive signal generation unit 53 obtains a deviation between the current command value I ^* and the actual current value I, and generates a motor drive signal so as to eliminate the deviation.

To be precise, the motor drive signal generator 53, a three-phase vector components of the current two-phase motor 31 by using the rotation angle theta _m, i.e. d-axis current and q-axis current in the d / q coordinate system Convert to a value. Then, the motor drive signal generation unit 53 calculates a deviation between the d-axis current value and the d-axis current command value and a deviation between the q-axis current value and the q-axis current command value, and calculates a PWM duty for eliminating these deviations. calculate. The motor drive signal generated by the motor drive signal generator 53 includes the PWM duty. When a current corresponding to the motor drive signal is supplied to the motor 31 through the inverter circuit 41, the motor 31 generates a rotational force according to the assist command value _Ta ^* .

Pinion angle computation section 54 takes in the rotation angle theta _m of the motor 31, calculates a pinion angle theta _p is the rotation angle of the pinion shaft 22c on the basis of the loaded rotation angle theta _m. As described above, the motor 31 is connected to the column shaft 22a via the speed reduction mechanism 32. Therefore, there is a correlation between the rotation angle theta _m and the pinion angle theta _p of the motor 31. It can be obtained pinion angle theta _p from the rotation angle theta _m of the motor 31 by utilizing the correlation. Further, as described above, the pinion shaft 22c is meshed with the rack shaft 23. Therefore, there is also a correlation between the amount of movement of the pinion angle theta _p and the rack shaft 23. That is, the pinion angle θ _p is a value that reflects the steered angle θ _ta of the steered wheels 26. The pinion angle θ _p is feedback-controlled based on a target pinion angle θ _p ^* described later.

<Assist command value calculation unit>
Next, the assist command value calculation unit 51 will be described in detail.
As shown in FIG. 2, the assist command value calculator 51 includes a basic assist component calculator 61, a target pinion angle calculator 62, and a pinion angle feedback controller (pinion angle F / B controller) 63. .

Basic assist component calculating unit 61 calculates a basic assist component _{T a1} ^* based on the vehicle speed V and the steering torque _{T h.} ^* Basic assist component _{T a1} is the base component of the assist command value _T ^{a *.} Basic assist component calculating unit 61 uses a three-dimensional map that defines the relationship between the steering torque T _h and the basic assist component T _a1 ^* in accordance with the vehicle speed V, the calculating a basic assist component T _a1 ^*. Basic assist component calculating unit 61, the larger the absolute value of the steering torque T _h is, or as the vehicle speed V is slow, and sets the absolute value of the basic assist component T _a1 ^* to a larger value.

Target pinion angle computing unit 62 takes in the basic assist component T _{a1 *} generated by the basic assist component calculating unit ^61, and the steering torque T _h, respectively. Target pinion angle calculation unit 62, when the sum of the basic assist component T _a1 ^* and the steering torque T _h a basic drive torque (input torque), have a ideal model to determine the ideal pinion angle based on the basic drive torque are doing. The ideal model is a model in which a pinion angle corresponding to an ideal steering angle corresponding to the basic drive torque is modeled in advance by an experiment or the like. Target pinion angle calculation unit 62 calculates the basic drive torque by adding the basic assist component T _a1 ^* and the steering torque T _h, computes the target pinion angle theta _p ^* on the basis of the ideal model from the basic drive torque thus determined I do. The target pinion angle calculation unit 62 takes in the vehicle speed V, the lateral acceleration LA, and the yaw rate YR, and takes into account the vehicle speed V, the lateral acceleration LA, and the yaw rate YR when calculating the target pinion angle θ _p ^* .

The pinion angle feedback control unit 63 takes in the target pinion angle θ _p ^* calculated by the target pinion angle calculation unit 62 and the actual pinion angle θ _p calculated by the pinion angle calculation unit 54, respectively. Pinion angle feedback control section 63, such that the actual pinion angle theta _p follows the target pinion angle θ _p ^*, PID (proportional, integral, derivative) control is carried out as a feedback control of the pinion angle. That is, the pinion angle feedback control section 63, a deviation between the actual pinion angle theta _p target pinion angle theta _p ^*, correction component T _{a2 *} ^{(compensation} component of the basic assist component to eliminate the deviation T _a1 ^* ).

Assist command value calculating section 51 calculates the assist command value _T ^{a *} by adding the correction component _{T a2} ^* to the basic assist component _{T a1} ^*.
<Target pinion angle calculator>
Next, the target pinion angle calculation unit 62 will be described in detail.

As described above, the target pinion angle computation unit 62 computes the target pinion angle theta _p ^* on the basis of the ideal model from the basic drive torque is the sum of the basic assist component T _a1 ^* and the steering torque T _h. The ideal model, the torque applied to the steering shaft 22, i.e., the basic drive torque T _p ^* described above, a model using by being represented by the following formula (A).

_{^{T p * = Jθ p * '}} ' + Cθ p * '+ Kθ p * ... (A)
Here, “J” is the moment of inertia of the steering wheel 21 and the steering shaft 22, “C” is a viscosity coefficient (friction coefficient) corresponding to the friction of the rack shaft 23 with the housing, and “K” is the steering wheel 21 and the steering shaft 22. Is a spring coefficient when each is regarded as a spring.

As it is seen from equation (A), the basic drive torque T _p ^*, the target pinion angle theta _p ^* of the second order time differential value θ _{p *} ^'' to a value obtained by multiplying the inertial moment J, the target pinion angle theta _p ^* one It is obtained by adding a value obtained by multiplying the differential time value θ _p ^* 'by the viscosity coefficient C and a value obtained by multiplying the target pinion angle θ _p ^* by the spring coefficient K.

The target pinion angle calculation unit 62 calculates the target pinion angle θ _p ^* according to the ideal model based on the equation (A).
As shown in FIG. 3, the ideal model based on the equation (A) is divided into an ideal EPS model 71 and an ideal vehicle model 72.

  The ideal EPS model 71 is tuned in accordance with the characteristics of each component of the electric power steering device 10, such as the steering shaft 22 and the motor 31. The ideal EPS model 71 has an adder 73, a subtractor 74, an inertia model 75, a first integrator 76, a second integrator 77, and a viscous model 78.

The adder 73 calculates the basic drive torque _T ^{p *} by adding the basic assist component _{T a1} ^* and the steering torque _{T h.}
The subtractor 74 subtracts a later-described viscous component T _vi ^* and a spring component T _sp ^* from the basic drive torque T _p ^* calculated by the adder 73. Here, the basic drive torque _T ^{p *} values viscous component _{T vi} ^* and the spring component _{T sp} ^* is subtracted and the subtraction value _T ^{p **.}

The inertia model 75 functions as an inertia control operation unit corresponding to the inertia term of the equation (A). The inertia model 75 calculates the pinion angular acceleration α _p ^* by multiplying the subtraction value T _p ^** calculated by the subtractor 74 by the reciprocal of the inertia moment J.

The first integrator 76 calculates the pinion angular velocity ω _p ^* by integrating the pinion angular acceleration α _p ^* calculated by the inertial model 75.
The second integrator 77 calculates the target pinion angle θ _p ^* by further integrating the pinion angular velocity ω _p ^* calculated by the first integrator 76. The target pinion angle θ _p ^* is an ideal rotation angle of the pinion shaft 22c based on the ideal EPS model 71.

The viscosity model 78 functions as a viscosity control calculation unit corresponding to the viscosity term in the equation (A). The viscosity model 78 calculates the viscosity component T _vi ^* of the basic drive torque T _p ^* by multiplying the pinion angular velocity ω _p ^* calculated by the first integrator 76 by the viscosity coefficient C.

The ideal vehicle model 72 is tuned according to the characteristics of the vehicle on which the electric power steering device 10 is mounted. The characteristics of the vehicle that affect the steering characteristics are determined, for example, by the specifications of the suspension and the wheel alignment, the grip force (friction force) of the steered wheels 26, 26, and the like. The ideal vehicle model 72 functions as a spring characteristic control calculation unit corresponding to the spring term in equation (A). The ideal vehicle model 72 calculates the spring component T _sp ^* of the basic drive torque T _p ^* by multiplying the target pinion angle θ _p ^* calculated by the second integrator 77 by the spring coefficient K. In calculating the spring component T _sp ^* , the ideal vehicle model 72 takes into account the vehicle speed V, the lateral acceleration LA, and the yaw rate YR.

According to the target pinion angle calculating section 62 thus constructed, the moment of inertia J and the viscosity coefficient C of the ideal EPS model 71, as well as by adjusting each spring coefficient K of the ideal vehicle model 72, the basic drive torque T _p ^* It is possible to directly tune the relationship between the target pinion angle θ _p ^* and thereby achieve a desired steering characteristic.

In this example, a target pinion angle θ _p ^* is set from the basic drive torque T _p ^* based on the ideal EPS model 71 and the ideal vehicle model 72, so that the actual pinion angle θ _p matches the target pinion angle θ _p ^*. Is feedback controlled. As described above, there is a correlation between the pinion angle theta _p and the steered angle theta _ta of the steered wheels 26, 26. Therefore, the steering operation of the steered wheels 26, 26 corresponding to the basic drive torque _T ^{p *} is also determined by the ideal EPS model 71 and the ideal vehicle model 72. That is, the steering feeling of the vehicle is determined by the ideal EPS model 71 and the ideal vehicle model 72. Therefore, a desired steering feeling can be realized by adjusting the ideal EPS model 71 and the ideal vehicle model 72.

Further, the actual turning angle θ _ta is maintained at the turning angle θ _ta corresponding to the target pinion angle θ _p ^* . Therefore, the effect of suppressing the reverse input vibration generated due to the road surface condition or the disturbance such as braking can be obtained. That is, in the case where the vibration to the steering mechanism 20 via the steering wheel 26 is transmitted, is adjusted correction component T _a2 ^* As pinion angle theta _p becomes the target pinion angle theta _p ^*. Therefore, the actual steering angle theta _ta is maintained to the turning angle theta _ta corresponding to the target pinion angle theta _p ^* defined by the ideal model. As a result, since the steering assist is performed in a direction to cancel the reverse input vibration, the transmission of the reverse input vibration to the steering wheel 21 is suppressed.

As described above, through the feedback control of the pinion angle θ _p, a sense of rigidity (so-called solid feeling) of the steering can be obtained. However, since the steering reaction force (the response felt through the steering), which is the force (torque) acting in the direction opposite to the steering direction of the driver, only corresponds to the target pinion angle θ _p ^* , It is concerned. For example, when the lateral acceleration acting on the vehicle increases due to the steering, the driver feels the lateral acceleration LA corresponding to the steering amount to the body, but the steering reaction force does not change, so that the sense of unity with the vehicle is insufficient depending on the driving situation. There is a risk. In addition, since the steering reaction force does not change depending on the road surface condition, it is difficult for the driver to grasp the road surface condition through the steering reaction force, and the driver may feel a sense of discomfort with the difference between the response imagined by the driver and the actual response. Therefore, in this example, the target pinion angle calculation unit 62, specifically, the ideal vehicle model 72 is configured based on the viewpoint of solving such a concern.

<Ideal vehicle model>
Next, the ideal vehicle model 72 will be described in detail.
As shown in FIG. 4, the ideal vehicle model 72 includes a first vehicle reaction force model 81, a second vehicle reaction force model 82, a first distribution gain calculator 83, a second distribution gain calculator 84, It has a value selector 85 and an interpolation calculator 86.

<First vehicle reaction force model>
The first vehicle reaction force model 81 calculates a first spring reaction force torque T _sp1 ^* . The first spring reaction torque T _sp1 ^* is a steering reaction force component (a reaction force component to be applied to the steering) according to the target pinion angle θ _p ^* , and contributes to a sense of rigidity (a sense of firmness) of the steering. . The first vehicle reaction force model 81 captures the target pinion angle θ _p ^* calculated by the second integrator 77, and multiplies the captured target pinion angle θ _p ^* by a spring coefficient K to obtain a target pinion angle θ _p ^*. The first spring reaction torque T _sp1 ^* according to θ _p ^* is calculated. The first spring reaction torque _Tsp1 ^* is a first reaction force component of the above-described spring component _Tsp ^* .

<Second vehicle reaction force model>
The second vehicle reaction force model 82 calculates a second spring reaction force torque T _sp2 ^* . The second spring reaction torque T _sp2 ^* is a steering reaction force component corresponding to the lateral acceleration LA acting on the vehicle, and has a characteristic close to a reaction force generated in an actual vehicle. The second spring reaction torque T _sp2 ^* contributes to a sense of unity with the vehicle.

The second spring reaction torque T _sp2 ^* is theoretically obtained by the following equation (B).
^{_{T sp2 * = (ζ / l}} n) · (I / l) · (l r · m · LA + I · γ ') ... (B)
However, "ζ" is trail amount, "l _n" is knuckle arm length, "I" is the yaw inertia moment acting on the vehicle, "l" is a wheel base, "l _r" is the front wheel axis when the vehicle is viewed from the side "M" is the weight of the vehicle, "LA" is the lateral acceleration acting on the vehicle, and "γ '" is the yaw angular acceleration, which is obtained by time-differentiating the yaw rate YR.

Here, the second spring reaction torque T _sp2 ^* can be obtained based on the equation (B), but in this example, the second spring reaction torque T _sp2 ^* is obtained by the following equation (C) in which “I · γ ′” in the equation (B) is omitted. 2 springs seek reaction torque _{T sp2} ^*. This is because the value of “I · γ ′” in equation (B) may be affected by noise.

^{_{T sp2 * = (ζ / l}} n) · (I / l) · (l r · m · LA) ... (C)
Therefore, the second vehicle reaction force model 82 can calculate the second spring reaction force torque T _sp2 ^* by applying the lateral acceleration LA acquired through the lateral acceleration sensor 440 to the equation (C). It is. The second spring reaction torque T _sp2 ^* is a second reaction force component of the above-described spring component T _sp ^* .

<First distribution gain calculator>
As shown in FIG. 4, the first distribution gain calculation unit 83 calculates the first distribution gain G1 using the first gain map 91 of the first distribution gain calculation unit 83. The first gain map 91 is a three-dimensional map that defines the relationship between the lateral acceleration LA and the first distribution gain G1 for each vehicle speed V (or vehicle speed range). The first distribution gain G1 is used to determine a use ratio between the first spring reaction torque _Tsp1 ^* and the second spring reaction torque _Tsp2 ^* .

  As shown in the graph of FIG. 5, when the lateral axis LA is plotted on the horizontal axis and the first distribution gain G1 is plotted on the vertical axis, the first gain map 91 has the following characteristics. That is, as the lateral acceleration LA increases and the vehicle speed V increases, the first distribution gain G1 is set to a larger value. Conversely, as the lateral acceleration LA decreases and the vehicle speed V decreases, the first distribution gain G1 is set to a smaller value. The first distribution gain G1 is a value in a range from “0” to “1”. Note that the first gain map 91 does not need to take the vehicle speed V into consideration.

<Second distribution gain calculation unit>
As shown in FIG. 4, the second distribution gain calculator 84 calculates a lateral acceleration difference value ΔLA. The lateral acceleration difference value ΔLA is obtained by the following equations (D) and (E).

ΔLA = | LA−LA ′ | (D)
Here, “LA” is the lateral acceleration detected by the lateral acceleration sensor 440, and “LA ′” is the estimated lateral acceleration based on the yaw rate YR and the vehicle speed V.

The estimated lateral acceleration LA 'is obtained by the following equation (E).
LA ′ = │YR · V│ (E)
Here, “YR” is the yaw rate detected by the yaw rate sensor 450, and “V” is the vehicle speed detected by the vehicle speed sensor 410.

The second distribution gain calculator 84 calculates the second distribution gain G2 using the second gain map 92 of the second distribution gain calculator 84 itself. The second gain map 92 is a three-dimensional map that defines the relationship between the lateral acceleration difference value ΔLA and the second distribution gain G2 for each vehicle speed V (or vehicle speed range). The second distribution gain G2 is also used to determine the usage ratio between the first spring reaction force torque _Tsp1 ^* and the second spring reaction force torque _Tsp2 ^* .

  As shown in the graph of FIG. 6, when the lateral acceleration difference value ΔLA is plotted on the horizontal axis and the second distribution gain G2 is plotted on the vertical axis, the second gain map 92 has the following characteristics. That is, as the lateral acceleration difference value ΔLA increases and the vehicle speed V increases, the second distribution gain G2 is set to a larger value. Conversely, the second distribution gain G2 is set to a smaller value as the lateral acceleration difference value ΔLA decreases and the vehicle speed V decreases. The second distribution gain G2 is a value in a range from “0” to “1”. Note that the second gain map 92 does not need to take the vehicle speed V into consideration.

<Maximum value selection section>
As shown in FIG. 4, the maximum value selection unit 85 takes in the first distribution gain G1 from the first distribution gain calculation unit 83 and the second distribution gain G2 from the second distribution gain calculation unit 84, and The larger one of the first distribution gain G1 and the second distribution gain G2 to be taken is selected.

<Interpolation calculation unit>
The interpolation calculation unit 86 uses the first distribution gain G1 or the second distribution gain G2 selected by the maximum value selection unit 85 to calculate the first spring reaction torque T _sp1 ^* and the second spring reaction force. determining the use ratio between the torque _{T sp2} ^*, calculates the basic drive torque _T ^{p *} of the spring component _{T sp} ^* based on the use ratio. The spring component T _sp ^* of the basic drive torque T _p ^* is obtained, for example, by the following equation (F).

T _sp ^* = T _sp1 ^* · (1-G) + T _sp2 ^* · G (F)
Here, “G” is a distribution gain, and the first distribution gain G1 or the second distribution gain G2 is applied.

In the above equation (F), the distribution gain G (the first distribution gain G1 or the second distribution gain G2) is set to a value from “0” to “1”. When the distribution gain G is “0”, the usage ratio of the first spring reaction force torque T _sp1 ^* becomes 100%. When the distribution gain G is “1”, the usage ratio of the second spring reaction torque T _sp2 ^* becomes 100%. When the distribution gain G is a value between “1” and “0”, the first spring reaction force torque T _sp1 ^* and the second spring reaction force torque T _sp2 ^* are the values of the distribution gain G, respectively. Are added at a predetermined usage ratio according to the above. In this way, the usage ratio of the first spring reaction torque T _sp1 ^* and the second spring reaction torque T _sp2 ^* is adjusted according to the value of the distribution gain G.

As shown in the graph of FIG. 5, as the lateral acceleration LA increases and the vehicle speed V increases, the first distribution gain G1 is set to a larger value. Therefore, as the lateral acceleration LA acting on the vehicle increases and the vehicle speed V increases, the usage ratio of the second spring reaction torque T _sp2 ^* increases, and the usage ratio of the first spring reaction torque T _sp1 ^* increases. Become smaller. Conversely, as the lateral acceleration LA decreases and the vehicle speed V decreases, the first distribution gain G1 is set to a smaller value. Therefore, as the lateral acceleration LA acting on the vehicle decreases and the vehicle speed V decreases, the usage rate of the second spring reaction torque T _sp2 ^* decreases, and the usage rate of the first spring reaction torque T _sp1 ^* decreases. growing.

As shown in the graph of FIG. 6, as the lateral acceleration difference value ΔLA increases and the vehicle speed V increases, the second distribution gain G2 is set to a larger value. For this reason, as the lateral acceleration difference value ΔLA increases and the vehicle speed V increases, the usage ratio of the second spring reaction torque T _sp2 ^* increases, and the usage ratio of the first spring reaction torque T _sp1 ^* decreases. Become. Conversely, as the lateral acceleration difference value ΔLA decreases and the vehicle speed V decreases, the first distribution gain G1 is set to a smaller value. Therefore, the smaller the lateral acceleration difference value ΔLA and the lower the vehicle speed V, the smaller the usage ratio of the second spring reaction torque T _sp2 ^{* and} the larger the usage ratio of the first spring reaction torque T _sp1 ^*. Become.

The interpolation calculation unit 86 adds the first spring reaction torque T _sp1 ^* and the second spring reaction torque T _sp2 ^* for which the usage ratios are set, as shown in the above equation (F). it allows calculating a basic drive torque _T ^{p *} of the spring component _{T sp} ^*.

<Operations and effects of the ideal vehicle model>
Next, the operation and effect of the ideal vehicle model 72 will be described. Here, there are two situations: a situation where the vehicle is traveling on a dry road (dry road) and a situation where the vehicle is traveling on a low-friction road (freezing road or snow-covered road) having a lower road surface friction resistance than the dry road. The description is divided into two driving situations.

<Dry path>
First, a case where the vehicle is traveling on a flat dry road will be described. In this case, since each tire of the vehicle grips the road surface, the actual lateral acceleration LA detected through the lateral acceleration sensor 440 and the estimated lateral acceleration LA ′ calculated by the above equation (E) are theoretically obtained. Matches. Therefore, the lateral acceleration difference value ΔLA calculated by the above equation (D) is theoretically “0” or a value equal to or less than the predetermined value ΔLA ₀ near “0” shown in the graph of FIG. Therefore, the maximum value selection unit 85 basically selects the first distribution gain G1.

As shown in the graph of FIG. 5, as the lateral acceleration LA increases, the first distribution gain G1 is set to a larger value. Therefore, as the lateral acceleration LA increases, the usage ratio of the second spring reaction torque T _sp2 ^* increases, and the usage ratio of the first spring reaction torque T _sp1 ^* decreases. Conversely, as the lateral acceleration LA decreases, the usage ratio of the second spring reaction torque T _sp2 ^* decreases, and the usage ratio of the first spring reaction torque T _sp1 ^* increases.

Here, the first spring reaction torque T _sp1 ^* based on the target pinion angle θ _p ^* contributes to a so-called steering rigidity (firm feeling). In addition, the second spring reaction torque _Tsp2 ^* based on the lateral acceleration LA contributes to a sense of unity with the vehicle. For this reason, by adjusting the usage ratio of the second spring reaction torque T _sp2 ^* and the first spring reaction torque T _sp1 ^* according to the lateral acceleration LA, it is possible to adjust the usage rate of the vehicle at that time. The feeling of rigidity of the steering and the feeling of unity with the vehicle can be suitably obtained.

For example, as the lateral acceleration LA increases, the sense of unity with the vehicle is required to be improved. However, as the lateral acceleration LA increases, the usage ratio of the second spring reaction torque _Tsp2 ^* increases, while the first ratio increases. The usage ratio of the spring reaction torque T _sp1 ^* is reduced. As the steering response increases as the lateral acceleration LA increases, a sense of unity with the vehicle according to the magnitude of the lateral acceleration LA can be obtained. In addition, since the lateral acceleration LA felt by the body matches the steering response, the driver can easily drive the vehicle.

Also, as the lateral acceleration LA decreases, the sense of unity with the vehicle does not become a problem. However, the use ratio of the second spring reaction torque _Tsp2 ^* decreases in accordance with the decrease in the lateral acceleration LA, while the first ratio increases. The usage ratio of the spring reaction torque T _sp1 ^* is increased. As a result, the sense of rigidity of the steering is easily transmitted to the driver's hand as a response, so that the driver can feel a sharp response through the steering.

As described above, the harmony between the sense of rigidity of the steering wheel and the sense of unity with the vehicle is achieved, so that the driver can easily drive the vehicle.
Further, as shown in the above equation (C), the second spring reaction torque T _sp2 ^* has a value corresponding to the value of the lateral acceleration LA. As can be seen from the equation (C), the second spring reaction force torque T _sp2 ^* can be regarded as a constant multiple of the lateral acceleration LA. Therefore, the value of the second spring reaction force torque T _sp2 ^* The value increases as the acceleration LA increases. As the second spring reaction force torque _{T sp2} ^* value increases, the subtraction value _T ^{p **,} inertia model 75 based on the thus the subtracted value _T ^{p **,} first integrator 76 and a second integrator The value of the target pinion angle θ _p ^* calculated through 77 decreases.

For correction component T _a2 ^* calculated by the pinion angle feedback control section 63 decreases according to the target pinion angle theta _p ^* decrease in. As the value of the correction component _Ta2 ^* added to the basic assist component _Ta1 ^* decreases, the magnitude of the assist command value _Ta ^* and, consequently, the value of the current command value I ^* decrease. Then, the motor torque of the motor 31 and, consequently, the steering assist force applied to the steering wheel decrease by the amount of the decrease in the current command value I ^* . That is, the steering reaction force felt by the driver through the steering increases. By increasing the steering reaction force in accordance with the magnitude of the lateral acceleration LA, it becomes easier for the lateral acceleration LA felt by the driver to feel in harmony with the response felt by the hand through the steering. Therefore, even when a large lateral acceleration LA is acting on the vehicle, the driver can easily feel a sense of unity with the vehicle.

<Low friction road>
Next, a case where the vehicle is traveling on a flat low friction road will be described. In this case, the actual lateral acceleration LA detected through the lateral acceleration sensor 440 and the estimated lateral acceleration LA ′ calculated by the above equation (E) are obtained because the grip of each tire on the road surface is reduced. Are basically inconsistent. For example, even when the running state of the vehicle is oversteer (including spin, etc.), the lateral acceleration LA hardly acts on the vehicle. The value is smaller than when traveling. Therefore, the first distribution gain G1 calculated through the first gain map 91 shown in FIG. 5 has a smaller value.

  On the other hand, the yaw rate YR is a value corresponding to the vehicle behavior. For this reason, the estimated lateral acceleration LA ′ calculated by the above equation (E) is a value that is more deviated from the actual lateral acceleration LA detected through the lateral acceleration sensor 440. Therefore, the lateral acceleration difference value ΔLA calculated by the above equation (D) becomes a larger value. Eventually, since the lateral acceleration difference value ΔLA becomes larger, the second distribution gain G2 calculated through the second gain map shown in FIG. 6 also becomes larger.

  As described above, the first distribution gain G1 tends to have a smaller value, whereas the second distribution gain G2 tends to have a larger value. For this reason, the maximum value selection section 85 basically selects the second distribution gain G2.

As shown in the graph of FIG. 6, the second distribution gain G2 is set to a larger value as the lateral acceleration difference value ΔLA increases (that is, as the vehicle slips). Therefore, as the lateral acceleration difference value ΔLA increases, the usage ratio of the second spring reaction torque T _sp2 ^* increases, and the usage ratio of the first spring reaction torque T _sp1 ^* decreases. Here, the first spring reaction torque T _sp1 ^* based on the target pinion angle θ _p ^* contributes to a so-called steering rigidity (firm feeling). In addition, the second spring reaction torque _Tsp2 ^* based on the lateral acceleration LA contributes to a sense of unity with the vehicle. For this reason, the use ratio of the second spring reaction torque T _sp2 ^* is set to be larger, and the use ratio of the first spring reaction torque T _sp1 ^* is set to be smaller. On the other hand, the sense of rigidity of the steering is further reduced. The driver can easily recognize that the vehicle is traveling on a low-friction road where it is difficult to feel a firm response as a steering feel by feeling the decrease in rigidity as a response. Since the response corresponding to the actual low friction road can be obtained, the driver can operate the steering wheel without discomfort.

Further, as described above, when the vehicle is traveling on a low friction road, lateral acceleration is hardly generated in the vehicle, so that the actually detected lateral acceleration LA becomes a smaller value. Here, as shown in the above equation (C), the second spring reaction force torque T _sp2 ^* has a value corresponding to the value of the lateral acceleration LA. As can be seen from the equation (C), the second spring reaction force torque T _sp2 ^* can be regarded as a constant multiple of the lateral acceleration LA. Therefore, the value of the second spring reaction force torque T _sp2 ^* The value decreases as the acceleration LA decreases. For this reason, the value of the spring component T _sp ^* of the basic drive torque T _p ^* calculated by the above equation (F) becomes a smaller value according to the value of the second spring reaction torque T _sp2 ^*. . Then, in the subtractor 74 of the ideal EPS model 71, a smaller value of the spring component T _sp ^* is subtracted from the basic drive torque T _p ^* . That is, the subtraction value T _p ^** calculated by the subtractor 74, the partial second spring reaction force torque T _sp2 ^* becomes smaller value, a larger value. Therefore, also increases in accordance with the subtraction value _T ^{p **} target pinion angle theta _p ^* value that is computed based on the subtraction value _T ^{p **.}

For correction component T _a2 ^* calculated by the pinion angle feedback control section 63 also increases according to the target pinion angle theta _p ^* increment of. As the value of the correction component _Ta2 ^* added to the basic assist component _Ta1 ^* increases, the magnitude of the assist command value _Ta ^* and, consequently, the value of the current command value I ^* increase. Then, the motor torque of the motor 31 and, consequently, the steering assist force applied to the steering wheel are increased by the amount of the current command value I ^* . That is, the steering reaction force felt by the driver through the steering is reduced. By feeling that the steering reaction force has decreased in response to the driver, the driver can more easily understand that the vehicle is traveling on a low friction road.

As shown in the graph of FIG. 6, in the present example, as the characteristics of the second gain map 92, as the lateral acceleration difference value ΔLA (absolute value) becomes larger starting from “0”, the vehicle speed V increases. Although the second distribution gain G2 is set to a larger value as much as possible, the second gain map 92 may have the following characteristics. That is, when the lateral acceleration difference value ΔLA (absolute value) is a value in the range of not less than “0” and not more than the predetermined value ΔLA ₀ as the threshold value, the value of the second distribution gain G2 is set to “0”. Is set. When the lateral acceleration differential value DerutaLA (absolute value) is a value exceeding a predetermined value DerutaLA _0, starting from the predetermined value DerutaLA _0, the lateral acceleration differential value DerutaLA (absolute value) becomes larger, and as the vehicle speed V becomes faster second Is set to a larger value. Here, the predetermined value ΔLA ₀ as a threshold value is a reference value for determining that the vehicle is traveling on a low friction road or that the vehicle is slipping, and is an experimental value. It is a value set in advance through, for example,.

  By providing the dead zone where the value of the second distribution gain G2 is "0" near "0" of the lateral acceleration difference value ΔLA, the following effects can be obtained. That is, as shown in the above equation (D), the lateral acceleration difference value ΔLA depends on the accuracy of the value of the estimated lateral acceleration LA ′. In particular, the vicinity of the lateral acceleration difference value ΔLA of “0” is easily affected by the accuracy of the value of the estimated lateral acceleration LA ′. For this reason, when the lateral acceleration difference value ΔLA is a value near “0”, the value of the second distribution gain G2 is set to “0” and the first distribution gain G1 is selected, so that an appropriate steering reaction force can be obtained. Higher reliability for the application can be ensured.

Incidentally, irrespective of whether the predetermined value ΔLA ₀ (dead zone) is set as described above, the usage ratio of the second spring reaction torque T _sp2 ^* to the increase in the lateral acceleration difference value ΔLA (the second distribution gain G2 The change characteristic of ()) may be appropriately changed depending on the required specifications and the like. For example, as the lateral acceleration difference value ΔLA increases, the second distribution gain G2 may increase linearly or stepwise. When the lateral acceleration difference value ΔLA increases, the usage ratio of the second spring reaction force torque T _sp2 ^* may be increased before the lateral acceleration difference value ΔLA increases.

<Second embodiment>
Next, a second embodiment in which the steering control device is embodied as an electric power steering device for a vehicle will be described. This embodiment has basically the same configuration as the first embodiment shown in FIGS. This example differs from the first embodiment in the method of calculating the spring component T _sp ^* by the ideal vehicle model 72.

As shown in FIG. 7, the assist command value calculation unit 51 has a rack axial force calculation unit 64. Rack axial force operation section 64 calculates the rack axial force T _af based on the steering torque T _h is detected through the three-phase phase current value I and the torque sensor 420, which is detected through the current sensor 45. Specifically, the rack shaft force calculating unit 64 calculates an assist torque generated by the motor 31 on the basis of the current value I, the rack shaft force by summing the assist torque that is the operation and the steering torque T _h Calculate T _af . The rack axial force T _af basically refers to a force applied in the axial direction of the rack shaft 23, and is calculated here as a torque for rotating the pinion shaft 22c. The rack shaft force has the same value Ideally the basic drive torque _T ^{p *} and (= basic assist component _{T a1} ^* + steering torque _{T h).}

As shown in FIG. 8, the ideal vehicle model 72, the target pinion angle theta _p ^*, the vehicle speed V, the addition to the lateral acceleration LA and yaw rate YR, further consideration of the rack shaft force _{T af} the basic drive torque _T ^{p *} spring _{Compute the} component T _sp ^* . The ideal vehicle model 72 has a third distribution gain calculator 87 and an interpolation calculator 88.

The third distribution gain calculator 87 calculates the third distribution gain G3 using the third gain map 93 of the third distribution gain calculator 87. The third gain map 93 is a three-dimensional map that defines the relationship between the lateral acceleration difference value ΔLA and the third distribution gain G3 for each vehicle speed V (or vehicle speed range). The third distribution gain G3 is a usage ratio of the spring component T _sp ^* of the basic drive torque T _p ^* calculated by the interpolation calculation unit 86 and the rack axial force T _af calculated by the rack axial force calculation unit 64. Used to determine For convenience of explanation, in the following description, the symbol “T _sp ^* ” of the spring component T _sp ^* calculated by the interpolation calculation unit 86 is written as the symbol “T _sp3 ^* ”, so that the final calculation is performed. Spring component T _sp ^* .

  As shown in the graph of FIG. 9, when the lateral acceleration difference value ΔLA is plotted on the horizontal axis and the third distribution gain G3 is plotted on the vertical axis, the third gain map 93 has the following characteristics. That is, the third distribution gain G3 is set to a larger value as the lateral acceleration difference value ΔLA increases and the vehicle speed V increases. Conversely, the third distribution gain G3 is set to a smaller value as the lateral acceleration difference value ΔLA decreases and the vehicle speed V decreases. The third distribution gain G3 is a value in a range from “0” to “1”. Note that the third gain map 93 does not need to take the vehicle speed V into consideration.

The interpolation calculation unit 88 uses the third distribution gain G3 calculated by the third distribution gain calculation unit 87 to calculate the spring component T _sp3 ^* calculated by the interpolation calculation unit 86 and the rack axial force calculation unit. the use ratio of the rack axial force T _af to be computed is determined by 64 and calculates the basic drive torque T _p ^* of the final spring component T _sp ^* based on the use ratio. The final spring component T _sp ^* is obtained by the following equation (G) based on the same concept as the above equation (F).

T _sp ^* = T _sp3 ^* · (1-G3) + T _af ^* · G3 (G)
In the equation (G), the third distribution gain G3 is set to a value from “0” to “1”. When the third distribution gain G3 is “0”, the usage ratio of the spring component T _sp3 ^* calculated by the interpolation calculation unit 86 becomes 100%. When the third distribution gain G3 is “1”, the usage ratio of the rack axial force T _af becomes 100%. When the third distribution gain G3 is a value between “1” and “0”, the spring component T _sp3 ^* and the rack axial force T _af calculated by the interpolation calculation unit 86 are equal to the third. Are added together at a predetermined usage ratio according to the value of the distribution gain G3 of. In this way, the use ratio between the spring component T _sp3 ^* calculated by the interpolation calculation unit 86 and the rack axial force T _af is adjusted according to the value of the third distribution gain G3.

According to the present embodiment, the following operations and effects can be obtained.
As described above, for example, when the vehicle is traveling on a low-friction road, it may be required to transmit the fact as road surface information to the driver as a steering reaction force. In this respect, according to this embodiment, when running on a low friction road, computing the basic drive torque T _p ^* of the final spring component T _sp ^* by adding a virtual rack axial force T _af Accordingly, it is possible to more appropriately transmit the road surface information to the driver as the steering reaction force.

As shown in the graph of FIG. 9, as the lateral acceleration difference value ΔLA increases, the third distribution gain G3 is set to a larger value. Therefore, the larger the lateral acceleration differential value DerutaLA, the use ratio of the rack axial force T _af is greater, the spring component T _sp3 ^* use ratio which is calculated by the previous interpolation operation unit 86 becomes smaller.

Here, the spring component T _sp3 ^* is calculated based on the viewpoint of harmonizing the sense of rigidity of the steering and the sense of unity with the vehicle according to the road surface condition or the vehicle behavior. The rack axial force _Taf is calculated as a driving force applied to the steering mechanism 20 (= steering torque _Th + assist torque based on the current value I). Further, as the road surface reaction force frictional resistance of the road becomes smaller smaller, as the road surface reaction force is reduced steering torque T _h is smaller, the rack shaft force T _af as the steering torque T _h is reduced becomes smaller. That is, the rack axial force T _af becomes smaller as the road surface reaction force becomes smaller. Conversely, the road surface reaction force frictional resistance of the road is increased larger, the more the steering torque T _h road surface reaction force is increased larger, the rack shaft force T _af as the steering torque T _h is increased becomes larger. That is, the rack axial force _Taf increases as the road surface reaction force increases. Thus, the rack axial force T _af reflects the road surface reaction force (reverse input load from the steered wheels 26 to the steering mechanism 20) although it is virtually.

Therefore, by setting the usage ratio of the rack axial force _Taf to be larger and the usage ratio of the spring component _Tsp3 ^* to be smaller, the final spring component _Tsp ^* calculated by the interpolation calculator 88 is Although it is virtually, the road surface condition is more strongly reflected. By using the final spring component T _sp ^{* in} which the road surface condition is more strongly reflected, it is possible to make the driver recognize the road surface condition as a more appropriate steering reaction force corresponding thereto.

For example, if the vehicle is traveling on a low friction road, the lateral acceleration differential value ΔLA a larger value as the road surface reaction force is small, and the rack shaft force T _af Where prone to a smaller value, the rack shaft force T The value of the final spring component T _sp ^* also becomes smaller according to _af . Then, as the value of the spring component T _sp ^* decreases, the value of the subtraction value T _p ^{** and} , consequently, the value of the target pinion angle θ _p ^* increase. Also, the minute the target pinion angle theta _p ^* increases, the correction component _{T a2} ^*, the assist command value _T ^{a *} and current command value ^{I *} increases. As the current command value I ^* increases, the motor torque of the motor 31 and, consequently, the steering assist force applied to the steering wheel increase. That is, the steering reaction force felt by the driver through the steering is reduced. By feeling that the steering reaction force has decreased in response to the driver, the driver can more easily understand that the vehicle is traveling on a low friction road.

Further, when the target pinion angle θ _p ^* is calculated using the rack axial force T _af as in this example, the following may be a concern. That is, the rack shaft force T _af, because is obtained by summing the assist torque and the steering torque T _h based on the current value I, which contains a viscosity and inertia of the steering mechanism 20. For this reason, the rack axial force _Taf does not reflect pure road surface information transmitted from the steered wheels 26 to the rack shaft 23 through the tie rods 25. Therefore, when calculating the target pinion angle θ _p ^* using the rack axial force T _af , there is a concern that the steering feeling may be reduced due to the influence of the viscosity and inertia of the steering mechanism 20.

In this regard, in this example, when calculating the target pinion angle θ _p ^* , the usage ratio of the rack axial force T _af in the final spring component T _sp ^* is adjusted according to the lateral acceleration difference value ΔLA. For example, when the vehicle is gripping on a flat road, the lateral acceleration difference value ΔLA becomes a smaller value. In this example, when calculating the target pinion angle θ _p ^* , the use ratio of the rack axial force T _af in the final spring component T _sp ^* is reduced according to the decrease in the lateral acceleration difference value ΔLA. Therefore, the influence of the viscosity and inertia of the steering mechanism 20 on the steering feeling is suppressed. Therefore, when the vehicle is gripping on a flat road, a more appropriate steering feel can be provided. As a result, road surface information appearing in the vehicle behavior (lateral acceleration difference value ΔLA) can be more appropriately transmitted to the driver as a steering reaction force.

In the present embodiment, the rack shaft force computing section 64 is computed the rack axial force T _af based on the three-phase phase current value I and the steering torque T _h, the rack shaft force based on the yaw rate YR and the lateral acceleration LA T _af may be calculated. Further, the target pinion angle calculation unit 62 may calculate the target pinion angle θ _p ^* using the basic drive torque T _p ^* instead of the rack axial force T _af calculated by the rack axial force calculation unit 64. . The basic driving torque T _p ^* is also a value reflecting the rack axial force. In this case, the interpolation calculation unit 88 determines the use ratio between the spring component T _sp3 ^* and the basic drive torque T _p ^* calculated by the interpolation calculation unit 86 using the third distribution gain G3, A final spring component T _sp ^* is calculated based on the usage ratio.

<Third embodiment>
Next, a third embodiment in which the steering control device is embodied as an electric power steering device for a vehicle will be described. This example is different from the second embodiment in the method of calculating the spring component T _sp ^* by the ideal vehicle model 72. In the second embodiment, the first spring reaction force torque T _sp1 ^* and the second spring G are used by using the distribution gain G having the larger value of the first distribution gain G1 and the second distribution gain G2. Although the usage ratio with the reaction torque T _sp2 ^* was adjusted, in this example, the use of the first spring reaction torque T _sp1 ^* and the second spring reaction torque T _sp2 ^* based on the first distribution gain G1. Adjust the ratio. In this example, the ideal vehicle model 72 is configured as follows.

  As shown in FIG. 10, the ideal vehicle model 72 includes a first vehicle reaction force model 81, a second vehicle reaction force model 82, a first distribution gain calculation unit 83, and two interpolation calculation units 86 and 88. are doing. That is, the ideal vehicle model 72 has a configuration in which the second distribution gain calculator 84 and the maximum value selector 85 shown in FIG. 8 are omitted. The third distribution gain calculator 87 calculates the lateral acceleration difference value ΔLA based on the above equations (D) and (E).

The interpolation calculation unit 86 determines the use ratio of the first spring reaction force torque T _sp1 ^* and the first spring reaction force torque T _sp1 ^* using the first distribution gain G1, and determines the use ratio. calculating a basic drive torque _T ^{p *} of the spring component _{T sp3} ^* based. The interpolation calculation unit 88 uses the third distribution gain G3 to use the spring component T _sp3 ^* calculated by the interpolation calculation unit 86 and the rack axial force T _af calculated by the rack axial force calculation unit 64. determines, calculates the basic drive torque T _p ^* of the final spring component T _sp ^* based on the use ratio.

Therefore, according to the present embodiment, the road surface condition can be more appropriately transmitted to the driver as the steering reaction force, as in the second embodiment. For example, when the vehicle is gripping on a flat dry road, the road surface condition can be transmitted to the driver through a firmer response based on a sense of rigidity. Further, when the vehicle is traveling on a low friction road, the road surface state can be transmitted to the driver through a steering reaction force in which the rack axial force T _af (road surface reaction force) is more strongly reflected.

Further, when the vehicle is traveling on a low friction road, the lateral acceleration LA is less likely to occur. As the lateral acceleration LA _decreases , the usage ratio of the first spring reaction torque T _sp1 ^* in the spring component T _sp3 ^* increases. Therefore, when the vehicle is traveling on a low friction road, the spring component _Tsp3 ^* tends to be mainly the first spring reaction force torque _Tsp1 ^* . When this spring component T _sp3 ^* is used as the final spring component T _sp ^* , the driver feels a higher sense of rigidity in response. For this reason, there is a possibility that the driver may feel uncomfortable with the fact that a firm response is obtained on a low friction road where it is originally difficult to feel a firm response as a steering feel.

In this respect, according to the present embodiment, using the ratio of the rack axial force with an increase in lateral acceleration differential value ΔLA T _af increases. Therefore, the final spring component T _sp ^* of the basic drive torque T _p ^* reflects the rack axial force T _af more. By using the final spring component _Tsp ^* , the road surface condition can be transmitted to the driver as a more appropriate steering reaction force corresponding to the road surface condition. For example, when the vehicle is traveling on a low friction road, the usage ratio of the spring component T _sp3 ^* mainly based on the first spring reaction torque T _sp1 ^* is further reduced, while the usage ratio of the rack axial force T _af is reduced. Is further increased. For this reason, the driver feels the steering reaction force according to the road surface reaction force as a response.

In the present embodiment, a second distribution gain calculator 84 may be provided instead of the first distribution gain calculator 83. In this case, the usage ratio between the first spring reaction torque T _sp1 ^* and the second spring reaction torque T _sp2 ^* is adjusted based on the second distribution gain G2.

<Fourth embodiment>
Next, a fourth embodiment in which the steering control device is embodied as an electric power steering device for a vehicle will be described. This embodiment differs from the third embodiment in that no consideration of the target pinion angle theta _{p *} ^(final spring component T _{sp ^*)} second spring reaction force torque T during operation of _sp2 ^*. In this example, the ideal vehicle model 72 is configured as follows.

As shown in FIG. 11, the ideal vehicle model 72 includes a first vehicle reaction force model 81, a fourth distribution gain calculator 89, and an interpolation calculator 90.
The fourth distribution gain calculator 89 calculates the lateral acceleration difference value ΔLA based on the above equations (D) and (E). The fourth distribution gain calculator 89 calculates a fourth distribution gain G4 using the fourth gain map 94 of the fourth distribution gain calculator 89. The fourth gain map 94 is a three-dimensional map that defines the relationship between the lateral acceleration difference value ΔLA and the fourth distribution gain G4 for each vehicle speed V (or vehicle speed range). The fourth distribution gain G4 is a usage ratio between the first spring reaction force torque T _sp1 ^* calculated by the first vehicle reaction force model 81 and the rack axial force T _af calculated by the rack axial force calculation unit 64. Used to determine

  As shown in the graph of FIG. 12, when the lateral acceleration difference value ΔLA is plotted on the horizontal axis and the fourth distribution gain G4 is plotted on the vertical axis, the fourth gain map 94 has the following characteristics. That is, the fourth distribution gain G4 is set to a larger value as the lateral acceleration difference value ΔLA increases and the vehicle speed V increases. Conversely, the fourth distribution gain G4 is set to a smaller value as the lateral acceleration difference value ΔLA decreases and the vehicle speed V decreases. The fourth distribution gain G4 is a value in a range from “0” to “1”. The fourth gain map 94 does not need to take the vehicle speed V into consideration.

The interpolation calculation unit 90 uses the fourth distribution gain G4 calculated by the fourth distribution gain calculation unit 89 to calculate the first spring reaction torque T _sp1 calculated by the first vehicle reaction force model 81. ^* and to determine the use ratio of the rack axial force T _af which is calculated by the rack axial force operation section 64 calculates the basic drive torque T _p ^* of the final spring component T _sp ^* based on the use ratio. The final spring component T _sp ^* is obtained by the following equation (H) based on the same concept as the above equation (F).

T _sp ^* = T _sp1 ^* · (1-G4) + T _af ^* · G4 (H)
In the equation (H), the fourth distribution gain G4 is set to a value from “0” to “1”. When the fourth distribution gain G4 is “0”, the usage ratio of the first spring reaction torque T _sp1 ^* becomes 100%. When the fourth distribution gain G4 is “1”, the usage ratio of the rack axial force T _af becomes 100%. When the fourth distribution gain G4 is a value between “1” and “0”, the first spring reaction torque T _sp1 ^* and the rack axial force T _af are respectively equal to the fourth distribution gain G4. The values are added at a predetermined usage ratio according to the value. In this way, the use ratio of the first spring reaction force torque T _sp1 ^* and the rack shaft force T _af is adjusted according to the value of the fourth distribution gain G4.

Therefore, according to the present embodiment, the road surface condition can be more appropriately transmitted to the driver as the steering reaction force, as in the third embodiment. For example, when the vehicle is a dry road grip running, since the lateral acceleration differential value ΔLA becomes a smaller value, the basic drive torque T _p ^* of the final spring component T _sp ^* first spring reaction torque T _sp1 ^* is mainly used. For this reason, the driver feels a sense of rigidity in response. Further, when the vehicle is traveling on a low friction road, the lateral acceleration difference value ΔLA becomes a larger value. Therefore, the final spring component T _sp ^* of the basic drive torque T _p ^* mainly includes the rack axial force T _af . And it will be. For this reason, the driver feels a smaller steering reaction force in response to the actual road surface condition as a response.

<Fifth embodiment>
Next, a fifth embodiment in which the steering control device is embodied as a steer-by-wire steering device for a vehicle will be described. The steer-by-wire refers to a steering device in which a steering wheel and a steered wheel are mechanically separated.

As shown in FIG. 13, the steering device 100 includes a reaction force motor 101, a rotation angle sensor 102, and a reaction force control unit 103 as a configuration for generating a steering reaction force. The reaction motor 101 is a source of a steering reaction force. The torque of the reaction motor 101 is applied to the steering shaft as a steering reaction force. The rotation angle sensor 102 detects the rotation angle θa of the reaction motor 101. Reaction force control unit 103 executes the reaction force control for generating the steering reaction force corresponding to the steering torque T _h through energization control of the reaction force motor 101.

  In addition, the steering device 100 includes a steering motor 104, a rotation angle sensor 105, and a steering control unit 106 as a configuration for generating a steering force that is power for steering a steered wheel. . The turning motor 104 is a source of the turning force. The torque of the turning motor 104 is applied to the turning shaft via a pinion shaft as a turning force. The rotation angle sensor 105 detects the rotation angle θb of the steering motor 104. The turning control unit 106 executes turning control for turning the steered wheels in accordance with the steering state through energization control of the turning motor 104.

  The reaction force control unit 103 includes a target steering reaction force calculation unit 111, a target steering angle calculation unit 112, a steering angle calculation unit 113, a steering angle feedback control unit 114, an adder 115, and an energization control unit 116.

Target steering reaction force calculation unit 111 calculates the target steering reaction force T ^* based on the steering torque T _h. Note that the target steering reaction force calculation unit 111 may calculate the target steering reaction force T ^* in consideration of the vehicle speed V. Target steering angle calculating unit 112 calculates the target steering angle theta ^* of the steering wheel based on the target steering reaction force T ^* and the steering torque T _h. The steering angle calculator 113 calculates the actual steering angle θs of the steering wheel based on the rotation angle θa of the reaction motor 101 detected through the rotation angle sensor 102. Steering angle feedback control section 114 calculates the steering angle correction amount T _[delta] of the actual steering angle θs through feedback control of the steering angle θs order to follow the target steering angle theta ^*. The adder 115 calculates the final target steering reaction force T ^* by adding the target steering reaction force T ^* a steering angle correction amount T _[delta]. The energization control unit 116 supplies power corresponding to the final target steering reaction force T ^* to the reaction motor 101. As a result, the reaction force motor 101 generates a torque corresponding to the final target steering reaction force T ^* . It is possible to give the driver an appropriate feeling of response according to the road surface reaction force.

The turning control unit 106 includes a pinion angle calculation unit 117, a pinion angle feedback control unit 118, and an energization control unit 119. Pinion angle calculating section 117 calculates the pinion angle theta _p is the actual rotation angle of the pinion shaft on the basis of the rotation angle θb of the steering motor 104 to be detected through the rotation angle sensor 105. The pinion angle feedback control unit 118 takes in the target steering angle θ ^* calculated by the target steering angle calculation unit 112 as the target pinion angle. Pinion angle feedback control section 118 calculates the pinion angle command value T _theta ^* through pinion angle theta _p feedback control so as to follow the actual pinion angle theta _p target steering angle theta ^*. The energization control unit 119 supplies power to the steering motor 104 according to the pinion angle command value _Tθ ^* . As a result, the steering motor 104 rotates by an angle corresponding to the pinion angle command value _Tθ ^* . That is, the steered wheels are steered by an amount corresponding to the operation of the steering wheel.

  Assuming such a configuration of the steering device 100, the calculation function of the target pinion angle calculation unit 62 in the first to fourth embodiments is added to the target steering angle calculation unit 112 and the calculation function of the pinion angle feedback control unit 63. Can be applied to the steering angle feedback control unit 114. Here, first, a case in which the first embodiment is applied to the steering device 100 will be described.

In this case, the target steering angle calculation unit 112 in the reaction force control unit 103 has the same functional configuration as the target pinion angle calculation unit 62 (see FIG. 2). While the previous target pinion angle calculation unit 62 captures the basic assist component _Ta1 ^* , the target steering angle calculation unit 112 of the present example captures the target steering reaction force T ^* . The fact that the target steering angle calculation unit 112 captures the steering torque _Th , the vehicle speed V, the yaw rate YR, and the lateral acceleration LA is the same as the previous target pinion angle calculation unit 62. The target pinion angle calculation unit 62 calculates the target pinion angle θ _p ^* , whereas the target steering angle calculation unit 112 of the present example calculates the target steering angle θ ^* . The content of the internal calculation processing of the target steering angle calculation unit 112 is the same as that of the target pinion angle calculation unit 62 described above, except that a part of the captured signal and the generated signal are different.

The steering angle feedback control unit 114 in the reaction force control unit 103 has the same function as the pinion angle feedback control unit 63 described above. The pinion angle feedback control unit 63 calculates the correction component T _a2 ^* for the basic assist component T _a1 ^* through the execution of the feedback control of the pinion angle θ _p , whereas the steering angle feedback control unit 114 of the present example uses the steering angle. The steering angle correction amount _Tδ is calculated by executing the feedback control of θs. Only the object of the feedback control is different, and the contents of the internal calculation processing of the steering angle feedback control unit 114 are the same as those of the pinion angle feedback control unit 63 described above.

  Now, in the steering device 100 of the steer-by-wire system, since the steering wheel and the steered wheels are mechanically separated, the road surface reaction force (rack axial force) from the tire may be directly transmitted to the steering wheel as the steering reaction force. Absent. For this reason, road surface information such as slipperiness of the road surface is hardly transmitted to the steering wheel in response.

In this regard, when the first embodiment is applied to the steering device 100, for example, when the vehicle is gripping on a flat dry road, the second component in the spring component _Tsp ^* is increased with an increase in the lateral acceleration LA. While the usage ratio of the spring reaction torque _Tsp2 ^* is increased, the usage ratio of the first spring reaction torque _Tsp1 ^* is reduced. Here, the first spring reaction torque T _sp1 ^* contributes to the sense of rigidity of the steering, and the second spring reaction torque T _sp2 ^* contributes to the sense of unity with the vehicle. Therefore, the target steering angle θ ^* more strongly reflects the second spring reaction force torque _Tsp2 ^* . Then, the reaction motor 101 generates a torque reflecting the target steering angle θ ^* as a steering reaction force. Accordingly, the sense of rigidity of the steering is further reduced in accordance with the increase in the lateral acceleration LA, while the sense of unity with the vehicle is further increased. A sense of unity with the vehicle according to the magnitude of the lateral acceleration LA can be obtained.

Further, when the vehicle is traveling on a low friction road such as a snowy road, the lateral acceleration LA is less likely to occur. Further, as the lateral acceleration difference value ΔLA increases, the usage ratio of the second spring reaction torque T _sp2 ^* in the spring component T _sp ^* increases, while the usage ratio of the first spring reaction torque T _sp1 ^* increases. Is reduced. For this reason, the first spring reaction torque T _sp1 ^* , which contributes to the sense of rigidity of the steering, is not easily reflected on the target steering angle θ ^* . The reaction motor 101 generates a torque reflecting the target steering angle θ ^* as a steering reaction force. By feeling the decrease in the sense of rigidity as a response, the driver can easily recognize that the vehicle is traveling on a low-friction road where it is originally difficult to feel a firm response as a steering feel.

Next, a case in which any one of the second to fourth embodiments is applied to the steering device 100 of the present example will be described.
In this case, as shown in FIG. 13, the reaction force control unit 103 is provided with a rack axial force calculation unit 64. The target steering angle calculator 112 calculates the target steering angle θ ^* in consideration of the rack axial force T _af calculated by the rack axial force calculator 64. Then, the vehicle when traveling a low friction road, the use ratio of the rack axial force T _af the spring component T _sp ^* with an increase of the lateral acceleration differential value ΔLA is increased. For this reason, the target steering angle θ ^* more strongly reflects the rack axial force _Taf . Therefore, the reaction force motor 101 generates a torque that more strongly reflects the rack axial force _Taf as a steering reaction force. For this reason, the driver easily recognizes that the vehicle is traveling on a slippery low friction road as the steering reaction force.

As described above, regardless of whether any of the first to fifth embodiments is applied to the steering device 100, the steering reaction force according to the rack axial force T _af (road surface reaction force) through the reaction force motor 101. Can be given to the steering wheel. Therefore, the road surface condition can be more appropriately transmitted to the driver as the steering reaction force.

<Other embodiments>
In addition, each embodiment may be modified and implemented as follows.
In the first to fourth embodiments, the torque sensor 420 is provided on the column shaft 22a, but may be provided on the intermediate shaft 22b or the pinion shaft 22c. If the steering torque T _h can be detected, it is possible to provide the appropriate position of the steering mechanism 20.

· In the first to fourth embodiments, the pinion angle feedback control section 63, but to perform the PID control on the pinion angle theta _p, may be carried out PI control. The same applies to the steering angle feedback control unit 114 in the fifth embodiment.

· In the first to fourth embodiments have to perform a feedback control for the pinion angle theta _p corresponding to the turning angle theta _ta of the steered wheels 26, 26, feedback control for the rotation angle of the intermediate shaft 22b May be performed. Further, feedback control may be performed on the rotation angle of the output shaft of the motor 31. Rotation angle of the output shaft of the intermediate shaft 22b and the motor 31, since both a value that reflects the turning angle theta _ta, through feedback control of the rotation angle, indirectly feedback control of the steering angle theta _ta It can be carried out. Alternatively, the steered angle θ _ta of the steered wheels 26, 26 may be detected, and direct feedback control may be performed on the steered angle θ _ta . In this case, the target pinion angle calculation section 62 functions as a target turning angle calculation section, and the pinion angle feedback control section 63 functions as a turning angle feedback control section. Similarly, in the fifth embodiment, feedback control of the steering angle θs is performed, but feedback control of the rotation angle θa of the reaction force motor 101 may be performed.

· In the first to fourth embodiments, the ideal EPS model 71 to determine the target pinion angle theta _p ^* (ideal pinion angle) based on the sum of the basic assist component _{T a1} ^* and the steering torque _{T h} was in may be obtained the target pinion angle theta _p ^* on the basis of only the steering torque T _h.

· In the first to fourth embodiments, the basic assist component calculating unit 61 has been to determine the basic assist component T _a1 ^* based on the steering torque T _h and the vehicle speed V, the basis of only the steering torque T _h The basic assist component _Ta1 ^* may be determined by using the following. Further, the basic assist component calculation unit 61 may execute at least one of the phase compensation control and the torque differential control. The phase compensation control, may change the phase of the steering torque T _h detected by the torque sensor 420 based on the assist gradient. In the torque differential control, it is desirable to increase the value of the basic assist component _Ta1 ^{* as} the differential value of the basic assist component _Ta1 ^* increases.

  In the first to fourth embodiments, the electric power steering device 10 that applies the steering assist force to the column shaft 22a has been described as an example, but for example, the electric power steering device 10 that applies the steering assist force to the pinion shaft 22c or the rack shaft 23. May be embodied in the electric power steering device.

DESCRIPTION OF SYMBOLS 10 ... Electric power steering apparatus, 20 ... Steering mechanism, 21 ... Steering wheel (rotating body), 22c ... Pinion shaft (rotating body), 23 ... Rack shaft (steering shaft), 26 ... Steering wheel, 31 ... Motor, 42 ... Microcomputer (steering control device), 61... Basic assist component calculating section (first calculating section), 62... Target pinion angle calculating section forming a second calculating section, 63... Forming a second calculating section. Pinion angle feedback control section, 440: lateral acceleration sensor, LA: lateral acceleration (actual lateral acceleration, state quantity reflecting road surface reaction force), LA ': estimated lateral acceleration, ΔLA: lateral acceleration difference value, I: current value , YR ... yaw rate (condition quantities road surface reaction force and the vehicle behavior is reflected), T _{a *} ^... assist command value, _{T a1 *} ^... basic assist component (basic component), _{T a2 *} ^... correction formed (Compensation component for base component), T h _... steering torque, T p _* ^... basic drive torque, T _{sp *} ^... spring component (the sum of the first and second components, the correction value for the basic component), T _sp1 ^* ... First corrected spring reaction torque (first reaction force component), T _sp2 ^* ... Second corrected spring reaction torque (second reaction force component and second component), T _af ^* ... Rack axial force (second component), T _sp3 ^* ... spring component (first component), V ... vehicle speed (state quantity reflecting vehicle behavior), θ _p ... pinion angle (actual rotation angle), θ _{p *} ^... target pinion angle (target rotation angle), θs ... steering angle (actual rotation angle), θ _{p *} ^... the target steering angle (target rotation angle).

Claims (6)

A steering control device that controls a motor that is a source of a driving force applied to a steering mechanism of a vehicle based on an assist command value calculated according to a steering operation,
A first calculation unit that calculates a basic component of the assist command value according to at least a steering torque;
A second calculation unit that calculates a compensation component for the basic component through feedback control that matches an actual rotation angle of the rotating body that rotates through the steering operation with a target rotation angle calculated according to the basic component,
The second calculation unit calculates at least a first component based on the target rotation angle and a second component based on a state quantity reflecting vehicle behavior or a road surface state, and calculates the first component and the Assuming that the target rotation angle is calculated in accordance with the base component corrected based on the added value with the second component,
The second arithmetic unit is configured to, when a lateral acceleration difference value that is a difference between an estimated lateral acceleration based on a state quantity reflecting a vehicle behavior or a road surface state and an actual lateral acceleration detected through an on-vehicle sensor increases, A steering control device having a calculation function part for increasing the use ratio of the second component before the acceleration difference value increases. The steering control device according to claim 1,
The steering control device, wherein the second calculation unit increases a usage ratio of the second component as the lateral acceleration difference value increases. In the steering control device according to claim 1 or 2,
The steering mechanism has a pinion shaft that rotates in conjunction with a steering operation, and a steered shaft that steers steered wheels in conjunction with the pinion shaft,
The steering control device, wherein the second component is a value of a current supplied to the motor and a rack axial force calculated based on the steering torque. In the steering control device according to any one of claims 1 to 3,
The second calculation unit calculates a first reaction force component based on the target rotation angle and a second reaction force component based on a yaw rate and a lateral acceleration, which are state quantities reflecting vehicle behavior, according to an actual lateral acceleration. Assuming that the first component is calculated by summing at the usage ratio determined by
The steering control device, wherein the second calculation unit increases a use ratio of the second reaction force component as the actual lateral acceleration increases. In the steering control device according to any one of claims 1 to 3,
The second calculation unit calculates a first reaction force component based on the target rotation angle and a second reaction force component based on a yaw rate and a lateral acceleration, which are state quantities reflecting vehicle behavior, as the actual lateral acceleration or the actual reaction. Assuming that the first component is calculated by summing at the use ratio determined according to the lateral acceleration difference value,
The second calculating unit is configured to increase the actual lateral acceleration or the lateral acceleration difference value as the actual lateral acceleration or the lateral acceleration difference value increases, of the increasing tendency of the actual lateral acceleration and the increasing tendency of the lateral acceleration difference value. A steering control device that increases the usage ratio of force components. In the steering control device according to claim 1 or 2,
The second calculating unit calculates a first reaction force component based on the target rotation angle as the first component, and also calculates a yaw rate and a lateral amount, which are state quantities reflecting vehicle behavior, as the second component. By calculating a second reaction force component based on the acceleration, and summing the first reaction force component and the second reaction force component at a use ratio determined according to the actual lateral acceleration or the lateral acceleration difference value, Assuming that a spring component as a correction value for the basic component is calculated,
The second calculating unit is configured to increase the actual lateral acceleration or the lateral acceleration difference value as the actual lateral acceleration or the lateral acceleration difference value increases, of the increasing tendency of the actual lateral acceleration and the increasing tendency of the lateral acceleration difference value. A steering control device that increases the usage ratio of force components.

Images