JP5265436B2 - Electric power steering apparatus and control method thereof - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a technology capable of restricting vibration in a control system and improving both stability and responding characteristics in the control system. <P>SOLUTION: The electric power steering device comprises a first rotary shaft connected to a steering wheel, a rack shaft for steering the steering wheel, a second rotary shaft for moving the rack shaft in a linear manner, a torsion bar, an electric motor for applying assisting power to the operation of the steering wheel, a steering torque detecting means for detecting the steering torque of the steering wheel and a desired current calculation part 20 for setting a desired current supplied to the electric motor according to the steering torque detected by the steering torque detecting means where the desired current calculation part 20 is operated in such a manner that a torsion bar is employed as a spring element, a resonance frequency compensation part 27 for restricting a resonance frequency component of a control system including the electric motor, the second rotary shaft and the rack shaft as its inertia elements is installed at an output side of the steering torque detecting means and then the desired current is set according to the steering torque having the resonance frequency component restricted by the resonance frequency compensation part 27. <P>COPYRIGHT: (C)2011,JPO&amp;INPIT

Description

  The present invention relates to an electric power steering apparatus and a control method thereof.

  In recent years, there has been proposed an electric power steering device that includes an electric motor in a steering system of a vehicle and assists a driver's steering force with the power of the electric motor. The control device that controls the electric power steering device determines a current to be supplied to the electric motor based on the detected steering torque in order to control the driving of the electric motor.

Then, when determining the current to be supplied to the electric motor based on the detected steering torque, the phase of the steering torque detected to increase the stability of the steering system is compensated by the phase compensator, and the phase compensated steering torque is There has been proposed a technique for supplying a current to an electric motor (see, for example, Patent Document 1).
Patent Document 2 proposes the following technique. That is, in view of the fact that oscillation occurs at the resonance frequency in the resonance system due to the spring element constituting the torque sensor and the handle inertia, a band elimination filter that removes the resonance frequency component of the signal of the motor drive control system on the output side of the torque sensor Is provided.

Japanese Patent Laid-Open No. 10-167086 JP-A-2-164665

Here, in the electric power steering apparatus, the electric motor is driven according to the steering torque applied to the steering wheel, and the driving force of the electric motor is transmitted to the pinion shaft constituting the pinion rack mechanism. Further, the driving force of the electric motor is transmitted to the rack shaft via the pinion shaft, and the direction of the steered wheels is changed by linearly moving the rack shaft.
Therefore, in order to suppress oscillation in the control system of the electric power steering apparatus, it is important to consider the electric motor, the pinion shaft, and the rack shaft as inertia elements.

  For this purpose, the present invention provides a first rotating shaft coupled to a steering wheel, a rack shaft for turning a steered wheel by linear movement, a second rotating shaft for linearly moving the rack shaft, A torsion bar that connects the first rotating shaft and the second rotating shaft and is twisted by the operation of the steering wheel, an electric motor that provides an assist force for the operation of the steering wheel, and steering of the steering wheel An electric power steering apparatus comprising: steering torque detection means for detecting torque; and target current setting means for setting a target current to be supplied to the electric motor based on the steering torque detected by the steering torque detection means. The target current setting means uses the torsion bar as a spring element, the electric motor, the second rotating shaft and the second rotating shaft. And a resonance compensation means for suppressing the resonance frequency component of the control system including the rack shaft as an inertia element on the output side of the steering torque detection means, and the steering torque whose resonance frequency component is suppressed by the resonance compensation means. The electric power steering apparatus is characterized in that the target current is set accordingly.

Here, it is preferable that the resonance compensation means has a filter function having an anti-resonance element of the control system and a low-pass filter function.
Moreover, it is preferable that the transfer function of the resonance compensating means includes the same element in the numerator as the denominator element of the transfer function of the control system.
In addition, it is preferable that the denominator of the transfer function of the resonance compensator has an order equal to or higher than the order of the numerator.

  From another point of view, the present invention provides a first rotating shaft coupled to the steering wheel, a rack shaft that steers the steered wheels by linear movement, and a second rotating shaft that linearly moves the rack shaft. An electric motor comprising: a torsion bar that connects the first rotating shaft and the second rotating shaft and is twisted by the operation of the steering wheel; and an electric motor that provides an assist force for the operation of the steering wheel. A method for controlling a power steering device, comprising: detecting a steering torque of the steering wheel; and using the torsion bar as a spring element in the detected steering torque, the electric motor, the second rotating shaft, and the rack shaft are inertial. The resonance frequency component of the control system included as an element is suppressed, and the electric power is controlled according to the steering torque in which the resonance frequency component is suppressed. It is a control method of an electric power steering apparatus characterized by setting the target current to be supplied to the motor.

  Here, when suppressing the resonance frequency component of the control system, it is preferable to use a filter function having an anti-resonance element and a low-pass filter function of the control system.

  According to the present invention, since the resonance frequency component of the control system including the torsion bar as a spring element and the electric motor, the second rotating shaft, and the rack shaft as inertia elements is suppressed, occurrence of oscillation can be suppressed with high accuracy. Can do. Thereby, stability in the control system can be improved.

It is a figure showing a schematic structure of an electric power steering device concerning a 1st embodiment. It is a schematic block diagram of the control apparatus of an electric power steering apparatus. It is a schematic block diagram of a target current calculation part. It is a schematic block diagram of a control part. It is a Bode diagram which compared the frequency characteristic of the control system with and without a resonance compensator. It is a figure which shows schematic structure of the electric power steering apparatus which concerns on 2nd Embodiment. It is a schematic block diagram of the control apparatus which concerns on 2nd Embodiment. It is a figure which shows schematic structure of the electric power steering apparatus which concerns on 3rd Embodiment. It is a schematic block diagram of the control apparatus which concerns on 3rd Embodiment. It is a figure which shows the relationship between correction coefficient (alpha) and the variation | change_quantity with respect to the reference value of the weight of a front wheel. It is a figure which shows the relationship between the correction coefficient (alpha) and a vehicle speed. It is a figure which shows the relationship between the correction coefficient (alpha) and the absolute value of a steering angle. It is a figure which shows the relationship between correction coefficient (beta) and the variation | change_quantity with respect to the reference value of the weight of a front wheel. It is a figure which shows the relationship between correction coefficient (beta) and vehicle speed. It is a figure which shows the relationship between correction coefficient (beta) and the absolute value of a steering angle.

Embodiments of the present invention will be described below in detail with reference to the accompanying drawings.
<First Embodiment>
FIG. 1 is a diagram showing a schematic configuration of an electric power steering apparatus 100 according to the first embodiment.
An electric power steering device 100 (hereinafter, also simply referred to as “steering device 100”) is a steering device for arbitrarily changing the traveling direction of a vehicle. In the present embodiment, a configuration applied to an automobile is exemplified. doing.

  The steering device 100 includes a wheel-like steering wheel (handle) 101 operated by a driver, and a steering shaft 102 provided integrally with the steering wheel 101. The steering shaft 102 and the upper connection shaft 103 are connected via a universal joint 103a, and the upper connection shaft 103 and the lower connection shaft 108 are connected via a universal joint 103b.

  The steering device 100 includes a tie rod 104 connected to each of the left and right front wheels 150 as steered wheels, and a rack shaft 105 connected to the tie rod 104. Further, the steering device 100 includes a pinion 106 a that constitutes a rack and pinion mechanism together with rack teeth 105 a formed on the rack shaft 105. The pinion 106 a is formed at the lower end portion of the pinion shaft 106.

  The steering device 100 also has a steering gear box 107 that houses the pinion shaft 106. The pinion shaft 106 is connected to the lower connection shaft 108 via a torsion bar (not shown) in the steering gear box 107. Inside the steering gear box 107, a torque sensor 109 is provided as an example of a steering torque detection unit that detects the steering torque of the steering wheel 101 based on the relative angle between the lower connecting shaft 108 and the pinion shaft 106.

The steering device 100 includes an electric motor 110 supported by the steering gear box 107, and a speed reducing mechanism 111 that decelerates the driving force of the electric motor 110 and transmits it to the pinion shaft 106.
In addition, the steering device 100 includes a motor current detection unit 33 (see FIG. 4) as an example of a current detection unit that detects the magnitude and direction of the actual current that actually flows through the electric motor 110, and the voltage between the terminals of the electric motor 110. Has a motor voltage detection unit 160 for detecting.

  The steering device 100 includes a control device 10 that controls the operation of the electric motor 110. The control device 10 receives the output value of the torque sensor 109, the output value of the vehicle speed sensor 170 that detects the vehicle speed of the vehicle, the output value of the motor current detection unit 33, and the output value of the motor voltage detection unit 160.

  The electric power steering apparatus 100 configured as described above detects the steering torque applied to the steering wheel 101 by the torque sensor 109, drives the electric motor 110 according to the detected torque, and generates the electric motor 110. Torque is transmitted to the pinion shaft 106. Thereby, the torque generated by the electric motor 110 assists the driver's steering force applied to the steering wheel 101.

Next, the control device 10 will be described.
The control device 10 is an arithmetic and logic circuit composed of a CPU, ROM, RAM, backup RAM, and the like.
FIG. 2 is a schematic configuration diagram of the control device 10 of the electric power steering device 100.
The control device 10 includes a torque signal Td in which the steering torque detected by the torque sensor 109 described above is converted into an output signal, and a vehicle speed signal v in which the vehicle speed detected by the vehicle speed sensor 170 is converted into an output signal. Is entered.

Further, the control device 10 converts the motor current signal Im obtained by converting the actual current detected by the motor current detector 33 into an output signal and the voltage detected by the motor voltage detector 160 into an output signal. The motor terminal voltage signal Vm is input.
Since the detection signal from the torque sensor 109 or the like is input as an analog signal, the control device 10 converts the analog signal into a digital signal by an A / D conversion unit (not shown) and takes it into the CPU.

  Then, the control device 10 calculates a target auxiliary torque based on the torque signal Td, a target current calculation unit 20 that calculates a target current necessary for the electric motor 110 to supply the target auxiliary torque, and a target current And a control unit 30 that performs feedback control and the like based on the target current calculated by the calculation unit 20.

Next, the target current calculation unit 20 will be described in detail. FIG. 3 is a schematic configuration diagram of the target current calculation unit 20.
The target current calculation unit 20 includes a base current calculation unit 21 that calculates a base current that serves as a reference for setting the target current, an inertia compensation current calculation unit 22 that calculates a current for canceling the inertia moment of the electric motor 110, and It has. Further, the target current calculation unit 20 calculates the rotation speed signal Nm of the electric motor 110 based on the damper compensation current calculation unit 23 that calculates the current that limits the rotation of the motor, the motor current signal Im, and the motor terminal voltage signal Vm. And a motor rotation speed estimation unit 24 for estimation.

Further, the target current calculation unit 20 includes a final target current determination unit 25 that determines a final target current based on outputs from the base current calculation unit 21, the inertia compensation current calculation unit 22, the damper compensation current calculation unit 23, and the like. I have.
Further, the target current calculation unit 20 performs phase compensation of the phase of the steering torque detected by the torque sensor 109, and resonance for removing the resonance frequency component of the steering torque phase-compensated by the phase compensation unit 26. And a resonance compensator 27 for performing compensation.

The phase compensation unit 26 performs a filtering process for phase compensation on the torque signal Td, which is an output value from the torque sensor 109, and outputs a torque signal Ts after the process. The resonance compensator 27 removes the resonance frequency component of the torque signal Ts and outputs the torque signal Tp from which the resonance frequency component has been removed. The resonance compensator 27 will be described in detail later.
The base current calculation unit 21 calculates a base current based on the steering torque detected by the torque sensor 109 and the vehicle speed detected by the vehicle speed sensor 170. More specifically, a base current is calculated based on a torque signal Tp that is an output value from the resonance compensator 27 and a vehicle speed signal v from the vehicle speed sensor 170, and a base current signal Ims that includes information on this base current. Is output. Note that the base current calculation unit 21, for example, maps the torque signal Tp and the vehicle speed to a map indicating the correspondence between the torque signal Tp and the vehicle speed signal v and the base current, which is previously created based on an empirical rule and stored in the ROM. The base current is calculated by substituting the signal v.

  The inertia compensation current calculation unit 22 calculates an inertia compensation current for canceling out the moment of inertia of the electric motor 110 and the system based on the torque signal Td and the vehicle speed signal v, and generates an inertia compensation current signal Is including information on this current. Output. For example, the inertia compensation current calculation unit 22 generates a torque signal Td on a map indicating the correspondence between the torque signal Td, the vehicle speed signal v, and the inertia compensation current, which is previously created based on an empirical rule and stored in the ROM. And the inertia compensation current is calculated by substituting the vehicle speed signal v.

The damper compensation current calculation unit 23 calculates a damper compensation current for limiting the rotation of the electric motor 110 based on the torque signal Td, the vehicle speed signal v, and the rotation speed signal Nm of the electric motor 110, and information on this current A damper compensation current signal Id including is output. The damper compensation current calculation unit 23 indicates the correspondence between the torque compensation signal Td, the vehicle speed signal v, the rotation speed signal Nm, and the damper compensation current, which are previously created based on empirical rules and stored in the ROM, for example. The damper compensation current is calculated by substituting the torque signal Td, the vehicle speed signal v, and the rotational speed signal Nm into the map.
The motor rotation speed estimation unit 24 estimates the rotation speed of the electric motor 110 based on the actual current detected by the motor current detection unit 33 and the voltage detected by the motor voltage detection unit 160. Details will be described later.

The final target current determination unit 25 includes the base current signal Ims output from the base current calculation unit 21, the inertia compensation current signal Is output from the inertia compensation current calculation unit 22, and the damper compensation output from the damper compensation current calculation unit 23. A final target current is determined based on the current signal Id, and a target current signal IT including information on this current is output. For example, the final target current determination unit 25 previously created a compensation current obtained by adding the inertia compensation current to the base current and subtracting the damper compensation current based on an empirical rule, and stored it in the ROM. The final target current is calculated by substituting it into a map indicating the correspondence between the compensation current and the final target current.
As described above, the target current calculation unit 20 functions as an example of a target current setting unit that sets a target current to be supplied to the electric motor 110 based on the steering torque detected by the torque sensor 109.

Next, the control unit 30 will be described in detail. FIG. 4 is a schematic configuration diagram of the control unit 30.
The control unit 30 includes a motor drive control unit 31 that controls the operation of the electric motor 110, a motor drive unit 32 that drives the electric motor 110, and a motor current detection unit 33 that detects the actual current that actually flows through the electric motor 110. have.

  The motor drive control unit 31 performs feedback control based on the deviation between the target current calculated by the target current calculation unit 20 and the actual current supplied to the electric motor 110 detected by the motor current detection unit 33. A feedback (F / B) control unit 40 and a PWM signal generation unit 60 that generates a PWM (pulse width modulation) signal for PWM driving the electric motor 110 are included.

The feedback control unit 40 includes a deviation calculating unit 41 for obtaining a deviation between the target current calculated by the target current calculating unit 20 and the actual current detected by the motor current detecting unit 33, and the deviation becomes zero. And a feedback (F / B) processing unit 42 for performing feedback processing.
The deviation calculation unit 41 outputs a deviation value between the output value IT from the target current calculation unit 20 and the output value Im from the motor current detection unit 33 as a deviation signal 41a.

The feedback (F / B) processing unit 42 performs feedback control so that the target current and the actual current match. For example, the input deviation signal 41a is proportionally processed by a proportional element. A signal obtained by the integration processing by the integration element is output, and the addition calculation unit adds these signals to generate and output a feedback processing signal 42a.
The PWM signal generation unit 60 generates the PWM signal 60a based on the output value from the feedback control unit 40, and outputs the generated PWM signal 60a.

  The motor drive unit 32 drives a motor drive circuit 70 in which four power field effect transistors are connected in an H-type bridge circuit configuration, and drives the gates of two field effect transistors selected from the four. And a gate drive circuit unit 80 for switching the field effect transistor. The gate drive circuit unit 80 selects two field effect transistors according to the steering direction of the steering wheel 101 based on the drive control signal (PWM signal) 60a output from the PWM signal generation unit 60, and selects the selected 2 The field effect transistors are switched.

  The motor current detection unit 33 detects the value of the motor current (armature current) flowing through the electric motor 110 from the voltage generated at both ends of the shunt resistor 71 connected in series with the motor drive circuit 70, and outputs the motor current signal Im. To do.

Next, the resonance compensator 27 will be described.
The steering device 100 is a control system including a torsion bar (not shown) used for detection of steering torque as a spring element, and an electric motor 110, a pinion shaft 106, and a rack shaft 105 as inertia elements. Therefore, for example, when the feedback processing unit 42 includes a proportional element and an integral element, if the values of the proportional gain and integral gain of the feedback processing unit 42 are increased to increase the response of the entire system, the control system The system tends to become unstable (oscillating) near the resonance frequency.

The resonance compensator 27 is provided to remove or suppress a peak in the resonance frequency band of the control system.
The transfer function H ₁ (s) indicating the characteristic of the resonance compensator 27 is defined as follows with respect to the transfer function G ₁ (s) indicating the characteristic of the control system of the steering device 100. That is, H ₁ (s) has the same element as the denominator of G ₁ (s) in the numerator, and the order of the denominator is a second order higher than the numerator in order to ensure the feasibility of the resonance compensator 27. And In other words, the resonance compensator 27 determines its transfer function so as to have a filter function having an anti-resonance element of the control system and a low-pass filter function.

First, consider the transfer function G ₁ (s).
The speed reduction mechanism 111 of the steering device 100 includes a worm wheel (not shown) attached to the pinion shaft 106 and a worm gear (not shown) attached to the output shaft of the electric motor 110.
In this case, the torque of the electric motor 110 is τ _m (N · m), the rotation angle is θ ₁ (rad), the torque of the worm gear is τ ₁ (N · m), and the motor shaft inertia is J ₁ (kg · m ² ). Then, the equation of motion for the electric motor 110 is expressed by the following equation (1).

ここで、モータ軸イナーシャＪ_１は、モータイナーシャをＪ_ｍ（ｋｇ・ｍ^２）、ウォームギヤのイナーシャをＪ_Ｔ１（ｋｇ・ｍ^２）とすると、Ｊ_１＝Ｊ_ｍ＋Ｊ_Ｔ１である。

Here, the motor shaft inertia J ₁ is J ₁ = J _m + J _T1 where J _m (kg · m ² ) is the motor inertia and J _T1 (kg · m ² ) is the inertia of the worm gear.

The rotation angle of the pinion shaft 106 is θ ₂ (rad), the torque of the worm wheel is τ ₂ (N · m), the torque of the pinion 106a is τ ₃ (N · m), and the pinion shaft inertia is J ₂ (kg · m ² ), and the spring constant of the torsion bar is k _tb (N · m / rad), the equation of motion for the pinion shaft 106 is expressed by the following equation (2).

ここで、ピニオン軸イナーシャをＪ_２は、ウォームホイールのイナーシャをＪ_Ｔ２（ｋｇ・ｍ^２）、ピニオン１０６ａのイナーシャをＪ_Ｔ３（ｋｇ・ｍ^２）とすると、Ｊ_２＝Ｊ_Ｔ２＋Ｊ_Ｔ３である。

Here, assuming that the pinion shaft inertia is J ₂ , the inertia of the worm wheel is J _T2 (kg · m ² ), and the inertia of the pinion 106a is J _T3 (kg · m ² ), J ₂ = J _T2 + J _T3 .

  Further, assuming that the displacement of the rack shaft 105 is x (m), the mass is m (kg), and the rotation radius of the pinion 106a is r (m), the equation of motion for the rack shaft 105 is expressed by the following equation (3). Is done.

ここで、ラック軸１０５の変位ｘは、以下の式（４）で表される。
ｘ＝ｒ・θ_２・・・（４）

Here, the displacement x of the rack shaft 105 is expressed by the following formula (4).
x = r · θ ₂ (4)

Further, when the worm reduction ratio is γ ₁ and the rack and pinion ratio is γ ₂ (m / rev), they are expressed by the following equations (5) and (6), respectively.
γ ₁ = θ ₁ / θ ₂ = τ ₂ / τ ₁ (5)
γ ₂ = 2 · π · r (6)

  Expression (7) can be derived from Expressions (3), (4), and (6).

式（１）、（２）、（５）より、式（８）を導き出せる。

Expression (8) can be derived from Expressions (1), (2), and (5).

式（７）に式（８）を代入して整理すると式（９）を導き出せる。

By substituting equation (8) into equation (7) and rearranging, equation (9) can be derived.

式（９）をラプラス変換して整理すると式（１０）を導き出せる。

When formula (9) is Laplace transformed and arranged, formula (10) can be derived.

なお、ｓはラプラス変換の演算子である。また、電動モータ１１０のトルクτ_ｍのラプラス変換をΤ_ｍ（ｓ）、ピニオンシャフト１０６の回転角度をθ_２のラプラス変換をΘ_２（ｓ）とする。
式（１０）により、上述した伝達関数Ｇ_１（ｓ）は式（１１）で表される。

Note that s is a Laplace transform operator. Further, the Laplace transform of the torque τ _m of the electric motor 110 is represented by Τ _m (s), and the rotation angle of the pinion shaft 106 is represented by θ ₂ as Θ ₂ (s).
From the equation (10), the above-described transfer function G ₁ (s) is represented by the equation (11).

また、式（１１）により、共振角周波数ω_１は式（１２）で表される。

Further, the resonance angular frequency ω ₁ is expressed by the equation (12) according to the equation (11).

したがって、共振補償部２７の特性を示す伝達関数Ｈ_１（ｓ）を「ａ_１・（（２πｆ_ｃ１）・（２πｆ_ｃ２））／（（ｓ＋２πｆ_ｃ１）・（ｓ＋２πｆ_ｃ２））」とする。なお、ａ_１は、式（１３）で表される値である。

Therefore, the transfer function H ₁ (s) indicating the characteristics of the resonance compensator 27 is defined as “a ₁ · ((2πf _c1 ) · (2πf _c2 )) / ((s + 2πf _c1 ) · (s + 2πf _c2 ))”. Incidentally, _{a 1} is a value represented by the formula (13).

そして、伝達関数Ｈ_１（ｓ）の分母の次数を共振補償部２７の実現性を確保する最低の次数である２次の次数とし、ローパスフィルタ（ＬＰＦ）の２段重ねとする。ｆ_ｃ１、ｆ_ｃ２はＬＰＦのカットオフ周波数である。

Then, the order of the denominator of the transfer function H ₁ (s) is set to the second order which is the lowest order that ensures the feasibility of the resonance compensator 27, and the low-pass filter (LPF) has two stages. f _c1 and f _c2 are cutoff frequencies of the LPF.

FIG. 5 is a Bode diagram comparing the frequency characteristics of the control system with and without the resonance compensator 27 being provided.
5A and 5B are diagrams showing the results of a simulation (numerical experiment) of the control system in which the operation of the steering wheel 101 is input and the rotation angle of the pinion 106a is output, where FIG. 5A is a gain characteristic diagram and FIG. FIG. In FIG. 5, the case where the resonance compensator 27 is provided is indicated by a solid line, and the case where the resonance compensator 27 is not provided is indicated by a broken line. Note that the denominator of the transfer function H ₁ (s) was set to a value corresponding to two stages of overlapping low-pass filters having a cutoff frequency of 100 (Hz).

  As shown in FIG. 5A, by providing the resonance compensator 27, the peak of the resonance frequency component can be reduced or canceled. Further, as shown in FIG. 5B, the phase lag can be greatly improved as compared with the case where the resonance compensator 27 is not provided. Thus, it can be seen that the stability is improved by providing the resonance compensator 27. Also, by providing the resonance compensator 27, the gain crossover frequency can be increased, and the responsiveness can also be improved.

Although the pinion type electric power steering apparatus 100 has been described so far, the column type electric power steering apparatus may also have a resonance compensator having a transfer function of H ₁ (s) as described above. Is preferred.

<Second Embodiment>
FIG. 6 is a diagram illustrating a schematic configuration of an electric power steering apparatus 200 according to the second embodiment.
Hereinafter, differences from the first embodiment will be described, and the same components will be denoted by the same reference numerals and detailed description thereof will be omitted.
The electric power steering apparatus 200 according to the second embodiment (hereinafter sometimes simply referred to as “steering apparatus 200”) is a so-called rack assist type electric power steering apparatus, and generates torque generated by the electric motor 201 in the rack. It is characterized in that it is applied to the shaft 105.

  That is, the electric motor 201 according to the second embodiment includes a stator (not shown) assembled to a housing (not shown) and a rack shaft 105 that can rotate with respect to the housing about the axis of the rack shaft 105 as a rotation center. And a rotor (not shown) assembled so as not to move in the axial direction. The rotor is engaged with a ball screw nut through an elastic body, and the rotor generates an assist force that rotates the ball screw nut through the elastic body to move the rack shaft 105 in the axial direction. The output is controlled by the control device 210 according to the second embodiment. Thus, the torque generated by the electric motor 201 is transmitted to the rack shaft 105 under the control of the control device 210, and the driver's steering force applied to the steering wheel 101 is assisted.

FIG. 7 is a schematic configuration diagram of the control device 210 according to the second embodiment.
Similar to the control device 10 according to the first embodiment, the control device 210 calculates a target auxiliary torque based on the torque signal Td, and the target current required for the electric motor 201 to supply the target auxiliary torque. And a control unit 230 that performs feedback control and the like based on the target current calculated by the target current calculation unit 220.
The control unit 230 has the same function and configuration as the control unit 30 of the control device 10 according to the first embodiment. The target current calculation unit 220 is different from the target current calculation unit 20 according to the first embodiment in that the target current calculation unit 220 includes a resonance compensation unit 271 that is different from the resonance compensation unit 27, and the others are the target according to the first embodiment. It has the same function and configuration as the current calculation unit 20.

The resonance compensator 271 is a resonance of a control system of the power steering apparatus 200 according to the second embodiment that includes a torsion bar (not shown) as a spring element and includes the electric motor 201, the pinion shaft 106, and the rack shaft 105 as inertia elements. It is provided to remove or suppress a peak near the frequency. Therefore, the transfer function H ₂ (s) indicating the characteristic of the resonance compensator 271 is defined as follows with respect to the transfer function G ₂ (s) indicating the characteristic of the control system of the steering device 200. That is, H ₂ (s) has the same element as the denominator of G ₂ (s) in the numerator, and the order of the denominator is the second order that is the lowest order that ensures the feasibility of the resonance compensator 271. , LPF two-tiered.

First, consider the transfer function G ₂ (s).
The torque of the electric motor 201 is τ _m (N · m), the rotation angle is θ ₁ (rad), the torque of the ball screw nut is τ ₁ (N · m), and the motor shaft inertia is J ₁ (kg · m ² ). Then, the equation of motion for the electric motor 201 is expressed by the following equation (14).

ここで、モータ軸イナーシャＪ_１は、モータイナーシャをＪ_ｍ（ｋｇ・ｍ^２）、ボールスクリューナットのイナーシャをＪ_Ｔ１（ｋｇ・ｍ^２）とすると、Ｊ_１＝Ｊ_ｍ＋Ｊ_Ｔ１である。

Here, the motor shaft inertia J ₁ is J ₁ = J _m + J _T1 where J _m (kg · m ² ) is the motor inertia and J _T1 (kg · m ² ) is the inertia of the ball screw nut.

Further, the torque of the pinion shaft 106 is τ ₂ (N · m), the rotation angle is θ ₂ (rad), the pinion shaft inertia is J ₂ (kg · m ² ), and the spring constant of the torsion bar is k _tb (N · m). / Rad), the equation of motion for the pinion shaft 106 is expressed by the following equation (15).

Further, if the displacement of the rack shaft 105 is x (m), the mass is m (kg), the rotation radius of the ball screw nut is r ₁ , and the rotation radius of the pinion 106a is r ₂ (m), the movement about the rack shaft 105 The equation is expressed by the following equation (16).

ここで、ラック軸１０５の変位ｘは、以下の式（１７）あるいは式（１８）で表される。
ｘ＝ｒ_１・θ_１＝γ_１／（２・π）×θ_１・・・（１７）
ｘ＝ｒ_２・θ_２＝γ_２／（２・π）×θ_２・・・（１８）
なお、γ_１は、ボールスクリューナットが１回転する間のラック軸１０５の移動距離を示すレシオ（ｍ／ｒｅｖ）であり、γ_１＝２・π・ｒ_１で表される。また、γ_２は、ピニオンシャフト１０６が１回転する間のラック軸１０５の移動距離を示すレシオ（ｍ／ｒｅｖ）であり、γ_２＝２・π・ｒ_２で表される。
また、式（１７）、（１８）より、式（１９）を導き出せる。
θ_１＝γ_２／γ_１×θ_２・・・（１９）
また、式（１６）、（１７）より、式（２０）を導き出せる。

Here, the displacement x of the rack shaft 105 is expressed by the following formula (17) or formula (18).
x = r ₁ · θ ₁ = γ ₁ / (2 · π) × θ ₁ (17)
x = r ₂ · θ ₂ = γ ₂ / (2 · π) × θ ₂ (18)
In addition, γ ₁ is a ratio (m / rev) indicating a moving distance of the rack shaft 105 during one rotation of the ball screw nut, and is expressed by γ ₁ = 2 · π · r ₁ . Γ ₂ is a ratio (m / rev) indicating the movement distance of the rack shaft 105 during one rotation of the pinion shaft 106, and is expressed by γ ₂ = 2 · π · r ₂ .
Further, equation (19) can be derived from equations (17) and (18).
θ ₁ = γ ₂ / γ ₁ × θ ₂ (19)
Further, equation (20) can be derived from equations (16) and (17).

そして、式（１４）〜（２０）より、式（２１）を導き出せる。

Then, Expression (21) can be derived from Expressions (14) to (20).

式（２１）をラプラス変換して整理すると式（２２）を導き出せる。

When formula (21) is Laplace transformed and arranged, formula (22) can be derived.

なお、ｓはラプラス変換の演算子である。また、電動モータ２０１のトルクτ_ｍのラプラス変換をΤ_ｍ（ｓ）、ピニオンシャフト１０６の回転角度をθ_２のラプラス変換をΘ_２（ｓ）とする。
式（２２）により、上述した伝達関数Ｇ_２（ｓ）は式（２３）で表される。

Note that s is a Laplace transform operator. Further, the Laplace transform of the torque τ _m of the electric motor 201 is Τ _m (s), and the Laplace transform of the pinion shaft 106 is θ ₂ is Θ ₂ (s).
From the equation (22), the above-described transfer function G ₂ (s) is represented by the equation (23).

また、式（２３）により、共振角周波数ω_２は式（２４）で表される。

Further, according to the equation (23), the resonance angular frequency ω ₂ is expressed by the equation (24).

Therefore, the order of the denominator of the transfer function H ₂ (s) indicating the characteristics of the resonance compensator 271 is set to the second order which is the lowest order to ensure the feasibility of the resonance compensator 271, and the LPF is two-tiered. . That is, “H ₂ (s) = a ₂ · ((2πf _c1 ) · (2πf _c2 )) / ((s + 2πf _c1 ) · (s + 2πf _c2 ))”. Incidentally, _{a 2} is the value represented by the formula (25). Further, f _c1 and f _c2 are cutoff frequencies of the LPF.

By providing the resonance compensator 271 whose transfer function is H ₂ (s), the peak of the resonance frequency component can be reduced or canceled also in the power steering device 200 according to the present embodiment. Further, the phase lag can be greatly improved as compared with the case where the resonance compensator 271 is not provided. Accordingly, the stability can be improved by providing the resonance compensator 271. Further, by providing the resonance compensator 271, the gain crossover frequency can be increased, so that responsiveness can also be improved.

<Third Embodiment>
FIG. 8 is a diagram illustrating a schematic configuration of an electric power steering apparatus 300 according to the third embodiment.
Hereinafter, differences from the first embodiment will be described, and the same components will be denoted by the same reference numerals and detailed description thereof will be omitted.
The electric power steering device 300 according to the third embodiment (hereinafter, sometimes simply referred to as “steering device 300”) is a so-called double pinion type electric power steering device, and the generated torque of the electric motor 301 is It is characterized in that it is applied to the rack shaft 105 via the pinion 302a of the second pinion shaft 302. As shown in FIG. 8, the second pinion shaft 302 is a member provided separately from the pinion shaft 106 connected to the steering wheel 101 via a torsion bar.

  As described above, the steering device 300 according to the third embodiment includes the second pinion shaft 302, and the worm wheel 303 attached to the second pinion shaft 302 and the output shaft of the electric motor 301. An attached worm gear (not shown) is connected. The output of the electric motor 301 is controlled by the control device 310 according to the third embodiment. Thus, the torque generated by the electric motor 301 is transmitted to the rack shaft 105 under the control of the control device 310, and the driver's steering force applied to the steering wheel 101 is assisted.

FIG. 9 is a schematic configuration diagram of a control device 310 according to the third embodiment.
Similar to the control device 10 according to the first embodiment, the control device 310 calculates the target auxiliary torque based on the torque signal Td, and the target current required for the electric motor 301 to supply the target auxiliary torque. The target current calculation unit 320 for calculating the target current and the control unit 330 for performing feedback control and the like based on the target current calculated by the target current calculation unit 320.
The control unit 330 has the same function and configuration as the control unit 30 of the control device 10 according to the first embodiment. The target current calculation unit 320 is different from the target current calculation unit 20 according to the first embodiment in that the target current calculation unit 320 includes a resonance compensation unit 272 that is different from the resonance compensation unit 27, and the others are the target according to the first embodiment. It has the same function and configuration as the current calculation unit 20.

The resonance compensator 272 uses the torsion bar (not shown) of the power steering apparatus 300 according to the third embodiment as a spring element, and the electric motor 301, the pinion shaft 106, the second pinion shaft 302, and the rack shaft 105 are inertial. It is provided to eliminate or suppress a peak in the vicinity of the resonance frequency of the control system included as an element. Therefore, the transfer function H ₃ (s) indicating the characteristic of the resonance compensator 272 is defined as follows with respect to the transfer function G ₃ (s) indicating the characteristic of the mechanical resonance system of the steering device 300. That is, H ₃ (s) has the same element as the denominator of G ₃ (s) in the numerator, and the order of the denominator is the second order that is the lowest order that ensures the feasibility of the resonance compensator 272. , LPF two-tiered.

First, consider the transfer function G ₃ (s).
If the torque of the electric motor 301 is τ _m (N · m), the rotation angle is θ ₁ (rad), the torque of the worm gear is τ ₁ (N · m), and the motor shaft inertia is J ₁ (kg · m ² ), The equation of motion for the electric motor 301 is expressed by the following equation (26).

ここで、モータ軸イナーシャＪ_１は、モータイナーシャをＪ_ｍ（ｋｇ・ｍ^２）、ウォームギヤのイナーシャをＪ_Ｔ１（ｋｇ・ｍ^２）とすると、Ｊ_１＝Ｊ_ｍ＋Ｊ_Ｔ１である。

Here, the motor shaft inertia J ₁ is J ₁ = J _m + J _T1 where J _m (kg · m ² ) is the motor inertia and J _T1 (kg · m ² ) is the inertia of the worm gear.

Further, the rotation angle of the second pinion shaft 302 is θ ₂ (rad), the torque of the worm wheel 303 is τ ₂ (N · m), and the torque of the pinion 302a of the second pinion shaft 302 is τ ₃ (N · m). ), Where the pinion shaft inertia is J ₂ (kg · m ² ), the equation of motion for the second pinion shaft 302 is expressed by the following equation (27).

ここで、ピニオン軸イナーシャをＪ_２は、ウォームホイールのイナーシャをＪ_Ｔ２（ｋｇ・ｍ^２）、第２のピニオンシャフト３０２のイナーシャをＪ_Ｔ３（ｋｇ・ｍ^２）とすると、Ｊ_２＝Ｊ_Ｔ２＋Ｊ_Ｔ３である。

Here, assuming that the pinion shaft inertia is J ₂ , the inertia of the worm wheel is J _T2 (kg · m ² ), and the inertia of the second pinion shaft 302 is J _T3 (kg · m ² ), J ₂ = J _T2 + J _T3 .

The rotation angle of the pinion shaft 106 is θ ₃ (rad), the torque of the pinion 106a of the pinion shaft 106 is τ ₄ (N · m), the pinion shaft inertia is J ₃ (kg · m ² ), and the spring constant of the torsion bar Is k _tb (N · m / rad), the equation of motion for the pinion shaft 106 is expressed by the following equation (28).

Further, the displacement of the rack shaft 105 is x (m), the mass is m (kg), the rotation radius of the pinion 302a of the second pinion shaft 302 is r ₁ (m), and the rotation radius of the pinion 106a of the pinion shaft 106 is r. Assuming ₂ (m), the equation of motion about the rack shaft 105 is expressed by the following equation (29).

ここで、ラック軸１０５の変位ｘは、以下の式（３０）で表される。
ｘ＝ｒ_１・θ_２＝ｒ_２・θ_３・・・（３０）
また、ウォーム減速比をγ_１、第２のピニオンシャフト３０２が１回転する間のラック軸１０５の移動距離を示すレシオをγ_２（ｍ／ｒｅｖ）、ピニオンシャフト１０６が１回転する間のラック軸１０５の移動距離を示すレシオをγ_３（ｍ／ｒｅｖ）とすると、それぞれ以下の式（３１）、（３２）、（３３）で表される。
γ_１＝θ_１／θ_２＝τ_２／τ_１・・・（３１）
γ_２＝２・π・ｒ_１・・・（３２）
γ_３＝２・π・ｒ_２・・・（３３）
式（２６）、（３０）、（３１）より、式（３４）を導き出せる。

Here, the displacement x of the rack shaft 105 is expressed by the following equation (30).
x = r ₁ · θ ₂ = r ₂ · θ ₃ (30)
Further, the worm reduction ratio is γ ₁ , the ratio indicating the movement distance of the rack shaft 105 during one rotation of the second pinion shaft 302 is γ ₂ (m / rev), and the rack shaft during one rotation of the pinion shaft 106. Assuming that the ratio indicating the moving distance of 105 is γ ₃ (m / rev), it is expressed by the following equations (31), (32), and (33), respectively.
γ ₁ = θ ₁ / θ ₂ = τ ₂ / τ ₁ (31)
γ ₂ = 2 · π · r ₁ (32)
γ ₃ = 2 · π · r ₂ (33)
Expression (34) can be derived from Expressions (26), (30), and (31).

また、式（２７）、（３０）より、式（３５）を導き出せる。

Further, Expression (35) can be derived from Expressions (27) and (30).

ゆえに、式（３４）、（３５）より、式（３６）を導き出せる。

Therefore, Expression (36) can be derived from Expressions (34) and (35).

また、式（２８）、（２９）、（３０）より、式（３７）を導き出せる。

Further, Expression (37) can be derived from Expressions (28), (29), and (30).

そして、式（３６）、（３７）より、式（３８）を導き出せる。

Then, Expression (38) can be derived from Expressions (36) and (37).

式（３８）に、式（３２）、（３３）を代入して得た式をラプラス変換して整理すると式（３９）を導き出せる。

Expression (39) can be derived by replacing the expression obtained by substituting Expressions (32) and (33) into Expression (38) by Laplace transform.

なお、ｓはラプラス変換の演算子である。また、電動モータ３０１のトルクτ_ｍのラプラス変換をΤ_ｍ（ｓ）、ピニオンシャフト１０６の回転角度をθ_３のラプラス変換をΘ_３（ｓ）とする。
式（３９）により、上述した伝達関数Ｇ_３（ｓ）は式（４０）で表される。

Note that s is a Laplace transform operator. Further, the Laplace conversion of the torque τ _m of the electric motor 301 is Τ _m (s), and the rotation angle of the pinion shaft 106 is θ _{3 is} the Laplace conversion Θ ₃ (s).
From the equation (39), the above-described transfer function G ₃ (s) is represented by the equation (40).

また、式（４０）により、共振角周波数ω_３は式（４１）で表される。

Further, according to the equation (40), the resonance angular frequency ω ₃ is expressed by the equation (41).

Then, the order of the denominator of the transfer function H ₃ (s) indicating the characteristics of the resonance compensator 272 is set to the second order that is the lowest order that ensures the feasibility of the resonance compensator 272, and the LPF is two-tiered. . That is, “H ₃ (s) = a ₃ · ((2πf _c1 ) · (2πf _c2 )) / ((s + 2πf _c1 ) · (s + 2πf _c2 ))”. Incidentally, _{a 3} is the value represented by the formula (42). Further, f _c1 and f _c2 are cutoff frequencies of the LPF.

Then, by providing the resonance compensator 272 whose transfer function is H ₃ (s), the peak of the resonance frequency component can be reduced or canceled also in the power steering device 300 according to the present embodiment. Further, the phase delay can be greatly improved as compared with the case where the resonance compensator 272 is not provided. Thus, the stability can be improved by providing the resonance compensator 272. In addition, by providing the resonance compensator 272, the gain crossover frequency can be increased, and the responsiveness can also be improved.

In addition, it is preferable to correct the transfer functions of the resonance compensators 27, 271, and 272 in the first to third embodiments described above using a correction coefficient that changes depending on the situation.
That is, the correction coefficient alpha, the transfer function _H n of the resonance compensator 27,271,272 in the first to third embodiments described above _{(s) = a n · (} (2πf c1) · (2πf c2)) _{/ ((s + 2πf c1)} · (s + 2πf c2)), the n = 1, 2, 3, the operator "s" in the Laplace transform in _{a n} by replacing the "alpha × s", the resonance compensator 27, It is preferable to use it as a transfer function of 271 and 272.

For example, the transfer function H ₁ (s) of the resonance compensator 27 in the first embodiment is corrected as shown in Expression (43).

  FIG. 10 is a diagram showing the relationship between the correction coefficient α and the amount of change with respect to the reference value of the weight of the front wheels 150 (a positive change amount when the weight is heavy, and a negative change amount when the weight is light). is there. For example, an optimal correction coefficient α corresponding to the amount of change in the weight of the front wheel 150 is derived in advance as shown in FIG. Then, the resonance compensators 27, 271, and 272 are maps prepared in advance and stored in the ROM that indicate the correspondence between the amount of change in the weight of the front wheel 150 and the correction coefficient α, or the change in the weight of the front wheel 150. The correction coefficient α is calculated by substituting the amount of change in the weight of the front wheel 150 into the relational expression between the amount and the correction coefficient α, and used for the transfer function.

When the weight of the front wheel 150 becomes heavier than the reference value, the resonance frequency of the control system including the front wheel 150 when the steering device 100 is mounted on the vehicle is ω _n (n = 1, 2, the resonance frequency becomes higher than ω _n when the weight of the front wheel 150 becomes lighter than the reference value. Therefore, as shown in FIG. 10, the correction coefficient α is 1 when the weight of the front wheel 150 is the reference value, decreases to 0.8 as the weight increases, and is heavier than a certain weight. In this case, 0.8 is set. Further, it increases to 1.2 as the weight of the front wheel 150 becomes lighter than the reference value, and is preferably set to 1.2 when the weight is lighter than a certain weight.

It is also preferable to change the correction coefficient α according to the vehicle speed. FIG. 11 is a diagram showing the relationship between the correction coefficient α and the vehicle speed. For example, an optimum correction coefficient α corresponding to the vehicle speed is derived in advance as shown in FIG. The resonance compensators 27, 271, and 272 are based on a map that indicates the correspondence between the vehicle speed signal v and the correction coefficient α, or a relational expression between the vehicle speed signal v and the correction coefficient α that is created in advance and stored in the ROM. Then, the correction coefficient α is calculated by substituting the vehicle speed signal v and used for the transfer function.
Since the resonance frequency is considered to decrease as the vehicle speed increases, as shown in FIG. 11, the correction coefficient α is 1 when the vehicle speed is zero, and decreases to 0.8 as the vehicle speed increases. When the vehicle speed is equal to or higher than a certain speed, 0.8 is preferable.

In addition, it is preferable to change the correction coefficient α according to the rotation angle (steering angle) of the steering wheel 101. FIG. 12 is a diagram showing the relationship between the correction coefficient α and the absolute value of the steering angle. For example, an optimal correction coefficient α corresponding to the absolute value of the rudder angle is derived in advance as shown in FIG. Then, the resonance compensators 27, 271, and 272 are maps prepared in advance and stored in the ROM that indicate the correspondence between the absolute value of the steering angle and the correction coefficient α, or the absolute value of the steering angle and the correction coefficient α. The correction coefficient α is calculated by substituting the detected steering angle into the relational expression, and used for the transfer function.
Since the resonance frequency is considered to decrease as the absolute value of the steering angle increases, the correction coefficient α is 1 when the absolute value of the steering angle is zero, as shown in FIG. As the value increases, the value decreases to 0.8. When the absolute value of the steering angle is greater than or equal to a certain value, 0.8 is preferable.

  Then, by using the resonance angular frequency corrected in accordance with the situation in the transfer function of the resonance compensators 27, 271, 272 as described above, the stability can be improved with higher accuracy and the responsiveness can also be improved. be able to.

Further, assuming that the correction coefficient is β, the transfer functions H _n (s) and (n = 1, 2, 3) of the resonance compensators 27, 271, 272 in the first to third embodiments described above are multiplied by β. It is preferable to use the corrected transfer function as each transfer function. That is, it is used by correcting as H _ng (s) = β × H _n (s), (n = 1, 2, 3).

  FIG. 13 is a diagram showing the relationship between the correction coefficient β and the amount of change with respect to the reference value of the weight of the front wheel 150 (a positive change amount when the weight is heavy, and a negative change amount when the weight is light). is there. For example, an optimum correction coefficient β corresponding to the amount of change in the weight of the front wheel 150 is previously derived based on empirical rules as shown in FIG. Then, the resonance compensators 27, 271, and 272 are maps prepared in advance and stored in the ROM that indicate the correspondence between the amount of change in the weight of the front wheel 150 and the correction coefficient β, or the change in the weight of the front wheel 150. The correction coefficient β is calculated by substituting the amount of change in the weight of the front wheel 150 into the relational expression between the amount and the correction coefficient β, and used for the transfer function.

  In addition, when the weight of the front wheel 150 is heavier than the reference value, the vibration when the steering device 100 is mounted on the vehicle is increased, and the weight of the front wheel 150 is lighter than the reference value. In some cases, the vibration is reduced. Therefore, as shown in FIG. 13, the correction coefficient β is 1 when the weight of the front wheel 150 is the reference value, increases to 1.2 as the weight increases, and is heavier than a certain weight. In this case, 1.2 is set. Further, it decreases to 0.8 as the weight of the front wheel 150 becomes lighter than the reference value, and is preferably set to 0.8 when the weight is lighter than a certain weight.

It is also preferable to change the correction coefficient β according to the vehicle speed. FIG. 14 is a diagram showing the relationship between the correction coefficient β and the vehicle speed. For example, an optimal correction coefficient β corresponding to the vehicle speed is derived in advance as shown in FIG. Then, the resonance compensators 27, 271, and 272 are based on a map indicating the correspondence between the vehicle speed signal v and the correction coefficient β, or a relational expression between the vehicle speed signal v and the correction coefficient β, which is created in advance and stored in the ROM. Then, the correction coefficient β is calculated by substituting the vehicle speed signal v and used for the transfer function.
Since the vibration should be suppressed as the vehicle speed becomes smaller, as shown in FIG. 14, the correction coefficient β is 1.2 when the vehicle speed is zero, and decreases to 1 as the vehicle speed increases. When the vehicle speed is higher than a certain speed, 1 is preferable.

In addition, it is preferable to change the correction coefficient β according to the rotation angle (steering angle) of the steering wheel 101. FIG. 15 is a diagram illustrating the relationship between the correction coefficient β and the absolute value of the steering angle. For example, an optimum correction coefficient β corresponding to the absolute value of the rudder angle is derived in advance as shown in FIG. Then, the resonance compensators 27, 271, and 272 are maps prepared in advance and stored in the ROM that indicate the correspondence between the absolute value of the steering angle and the correction coefficient β, or the absolute value of the steering angle and the correction coefficient β. The correction coefficient β is calculated by substituting the detected steering angle into the relational expression, and used for the transfer function.
Since the vibration should be suppressed as the absolute value of the rudder angle is smaller, the correction coefficient β is 1.2 when the absolute value of the rudder angle is zero, as shown in FIG. It decreases to 1 as the absolute value of becomes larger, and is preferably 1 when the absolute value of the steering angle is greater than or equal to a certain value.

It is also preferable to use both the correction coefficients α and β described above. That is, the transfer functions H _n (s) = a _n · ((2πf _c1 ) · (2πf _c2 )) / ((s + 2πf _c1 ) of the resonance compensators 27, 271, and 272 in the first to third embodiments described above. _{· (s + 2πf c2))} , the n = 1, 2, 3, the operator "s" in the Laplace transform in _{a n} with multiplying the β of _{a n} by replacing the "alpha × s", the resonance compensator 27 , 271 and 272 are preferably used as transfer functions.

  Also in such a case, the resonance compensators 27, 271, and 272 are maps prepared in advance and stored in the ROM that indicate the correspondence between the amount of change in the weight of the front wheel 150 and the correction coefficients α and β, or the front wheel 150. The correction coefficients α and β are calculated by substituting the weight change amount of the front wheel 150 into the relational expression between the weight change amount and the correction coefficients α and β, and used for the transfer function. It is preferable that the correction coefficients α and β have the relationships shown in FIGS. 10 and 13, respectively.

The resonance compensators 27, 271, and 272 are maps prepared in advance and stored in the ROM that indicate the correspondence between the vehicle speed signal v and the correction coefficients α and β, or the vehicle speed signal v and the correction coefficients α and β. By substituting the vehicle speed signal v into the relational expression, the correction coefficients α and β are calculated and used for the transfer function. It is preferable that the correction coefficients α and β have the relationships shown in FIGS. 11 and 14, respectively.
The resonance compensators 27, 271, and 272 are maps prepared in advance and stored in the ROM that indicate the correspondence between the absolute value of the steering angle and the correction coefficients α and β, or the absolute value of the steering angle and the correction coefficient. The correction coefficients α and β are calculated by substituting the detected steering angle into the relational expression with α and β, and used for the transfer function. The correction coefficients α and β preferably have the relationships shown in FIGS. 12 and 15, respectively.

  In the first, second, and third embodiments described above, the resonance compensators 27, 271, and 272 are the resonance frequency components of the torque signal Ts that is a signal after phase compensation by the phase compensator 26. Is described, and the torque signal Tp from which the resonance frequency component is removed has been described. However, the configuration is not particularly limited to this configuration.

  For example, the resonance compensators 27, 271, and 272 remove the resonance frequency component of the torque signal Td that is the output value from the torque sensor 109 and output the torque signal Tp, and the phase compensation unit 26 removes the resonance frequency component. The torque signal Tp may be subjected to filtering processing for phase compensation to output the torque signal Ts. In this case, the base current calculation unit 21 calculates the base current based on the torque signal Ts that is the output value from the phase compensation unit 26 and the vehicle speed signal v from the vehicle speed sensor 170.

DESCRIPTION OF SYMBOLS 10,210,310 ... Control apparatus 20,220,320 ... Target current calculation part 26 ... Phase compensation part 27,271,272 ... Resonance compensation part 30,230,330 ... Control part 33 ... Motor current detection 40, feedback control unit, 100, 200, 300 ... electric power steering device, 101 ... steering wheel, 102 ... steering shaft, 109 ... torque sensor, 110, 201, 301 ... electric motor

Claims (6)

A first rotating shaft coupled to the steering wheel;
A rack shaft that steers the steered wheels by linear movement;
A second rotating shaft for linearly moving the rack shaft;
A torsion bar that connects the first rotating shaft and the second rotating shaft and is twisted by operation of the steering wheel;
An electric motor for providing an assisting force for the operation of the steering wheel;
Steering torque detection means for detecting steering torque of the steering wheel;
Target current setting means for setting a target current to be supplied to the electric motor based on the steering torque detected by the steering torque detection means;
In the electric power steering apparatus with
The target current setting means includes
Resonance compensation means for suppressing a resonance frequency component of a control system including the torsion bar as a spring element and the electric motor, the second rotating shaft and the rack shaft as inertia elements is provided on the output side of the steering torque detecting means. And
The electric power steering apparatus according to claim 1, wherein the target current is set according to a steering torque whose resonance frequency component is suppressed by the resonance compensator.   2. The electric power steering apparatus according to claim 1, wherein the resonance compensation means has a filter function having an anti-resonance element of the control system and a low-pass filter function.   3. The electric power steering apparatus according to claim 1, wherein the transfer function of the resonance compensator includes in the numerator the same element as the denominator of the transfer function of the control system. 4.   The electric power steering apparatus according to claim 3, wherein a denominator of a transfer function of the resonance compensating means has an order equal to or greater than the order of the numerator. A first rotating shaft coupled to the steering wheel;
A rack shaft that steers the steered wheels by linear movement;
A second rotating shaft for linearly moving the rack shaft;
A torsion bar that connects the first rotating shaft and the second rotating shaft and is twisted by operation of the steering wheel;
An electric motor for providing an assisting force for the operation of the steering wheel;
A method for controlling an electric power steering apparatus comprising:
Detecting the steering torque of the steering wheel;
In the detected steering torque, the torsion bar as a spring element, the resonance frequency component of the control system including the electric motor, the second rotating shaft and the rack shaft as inertia elements is suppressed,
A control method for an electric power steering apparatus, wherein a target current to be supplied to the electric motor is set according to a steering torque in which a resonance frequency component is suppressed.   6. The method for controlling an electric power steering apparatus according to claim 5, wherein when the resonance frequency component of the control system is suppressed, a filter function having an anti-resonance element of the control system and a low-pass filter function are used. .

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