WO2014054657A1

WO2014054657A1 - Control device for electric vehicle and control method for electric vehicle

Info

Publication number: WO2014054657A1
Application number: PCT/JP2013/076743
Authority: WO
Inventors: 弘征小松; 翔大野; 雄史勝又
Original assignee: 日産自動車株式会社
Priority date: 2012-10-04
Filing date: 2013-10-01
Publication date: 2014-04-10
Also published as: JP5850171B2; JPWO2014054657A1

Abstract

In a control device for an electric vehicle which sets a motor torque command value on the basis of vehicle information and controls the torque of a motor connected to a driving wheel, the control device applies a filter processing to the motor torque command to reduce natural vibration frequency component in a back and forth direction under a spring and controls the motor torque in accordance with a final torque command value acquired by applying the filter processing to the motor torque command value.

Description

Electric vehicle control device and electric vehicle control method

The present invention relates to an electric vehicle control device and an electric vehicle control method.

In the vibration suppression control device for a vehicle described in JP2003-9566A, a torque target value for controlling a motor is calculated by the following method. First, a drive torque target value is calculated by performing a filtering process for removing or reducing the natural vibration frequency component of the vehicle torque transmission system on the drive torque request value of the drive motor calculated from the accelerator opening and the vehicle speed. . Subsequently, the motor rotation speed estimation value is calculated from the drive torque target value in consideration of the motor characteristic model, and the deviation between the calculated motor rotation speed estimation value and the actual motor rotation speed is calculated as the natural vibration frequency of the driving force transmission system. A torque command value is calculated by passing through a filter constituted by an inverse system of a band pass filter and a motor characteristic model as a center frequency. Then, the final drive torque target value is calculated by adding the calculated torque command value to the drive torque target value. This eliminates the effects of road gradients, torque transmission system disturbances, motor characteristic model errors, etc., and also eliminates or reduces the natural vibration frequency component of the vehicle torque transmission system, thereby reducing the damping effect and steep torque. Can be achieved at the same time.

By the way, the natural frequency of torsional vibration of a general vehicle drive shaft is about 6 to 10 Hz. Since the drive shaft is twisted with respect to the output torque, vibration is excited, and the vibration is expressed in the motor rotation speed. Therefore, the drive shaft can be removed using the technique described in JP2003-9566A. As a result, the output shaft torque of the vehicle can be raised sharply without causing overshoot or hunting and so as to reach a steady value in about 50 ms from the input of the torque command.

However, as a result of steeply raising the output shaft torque of the vehicle, a vibration component having a natural frequency of about 20 to 40 Hz may be excited. Vibrations with a natural frequency of about 20 to 40 Hz are excited when the vehicle accelerates and the reaction force of the driving force excites the motor unit (unsprung) including the suspension member back and forth via the lower arm. Therefore, no vibration component is generated in the drive shaft torque and the motor rotation speed. In other words, the vehicle vibration damping control device described in JP2003-9566A cannot remove unsprung longitudinal vibrations having a natural frequency of about 20 to 40 Hz.

The present invention aims to suppress vibration in the front-rear direction under the spring.

The control device for an electric vehicle in one embodiment sets a motor torque command value based on the vehicle information, and controls the torque of the motor connected to the drive wheels. The control device for an electric vehicle includes a filtering unit that applies a filtering process to the motor torque command value to reduce a natural vibration frequency component in the unsprung direction, and the filtering process is performed on the motor torque command value. Motor torque control means for controlling the motor torque in accordance with the final torque command value obtained by

Embodiments of the present invention will be described in detail below with reference to the accompanying drawings.

FIG. 1 is a block diagram illustrating a main configuration of an electric vehicle including a control device for an electric vehicle according to the first embodiment. FIG. 2 is a flowchart showing a flow of processing of motor current control performed by the electric motor controller. FIG. 3 is a diagram showing an example of an accelerator opening-torque table. FIG. 4A is a perspective view showing an unsprung structure on the lower side of the suspension spring. FIG. 4B is a top view showing an unsprung structure on the lower side of the suspension spring. FIG. 5 is a diagram illustrating a configuration of the vehicle longitudinal vibration suppression filter. FIG. 6 is an example of a control block diagram for performing processing for calculating the third torque command value Tm3 *. FIG. 7 is an example of a control block diagram illustrating the configuration of the vibration suppression control FF calculation unit 402B. FIG. 8 is another example of a control block diagram illustrating the configuration of the vibration suppression control FF calculation unit 402B. FIG. 9 is a detailed block diagram of the vibration suppression control FB calculation unit 600A. FIG. 10 is a diagram illustrating the characteristics of the transfer function H (s). FIG. 11 is a configuration diagram of an unsprung spring / mass / damper model. FIG. 12A is a diagram for explaining the response of acceleration when the torque command value is increased in steps at the time of sudden acceleration from the creep state, and is a control device for an electric vehicle in the first and second embodiments. It is a figure which shows the control result by. FIG. 12B is a diagram for explaining the response of acceleration when the torque command value is increased stepwise at the time of sudden acceleration from the creep state, and an example of a control result by the control device described in JP2013-9566A FIG. FIG. 13A is a diagram for explaining the response of acceleration when the torque command value is increased stepwise during sudden acceleration from the coast state, and is based on the control device for the electric vehicle in the first and second embodiments. It is another figure for demonstrating a control result. FIG. 13B is a diagram for explaining an acceleration response when the torque command value is increased stepwise during sudden acceleration from the coast state, and shows an example of a control result by the control device described in JP2013-9566A. FIG.

-First embodiment-
FIG. 1 is a block diagram illustrating a main configuration of an electric vehicle including a control device for an electric vehicle according to the first embodiment. The control apparatus for an electric vehicle according to the present invention includes an electric motor as a part or all of the drive source of the vehicle, and can be applied to an electric vehicle that can be driven by the driving force of the electric motor. And can be applied to fuel cell vehicles.

The electric motor controller 2 uses, as digital signals, signals indicating the vehicle state such as the vehicle speed V, the accelerator pedal opening AP, the rotor phase α of the electric motor (three-phase AC motor) 4, and the currents iu, iv, iw of the electric motor 4. Based on the input signal, a PWM signal for controlling the electric motor 4 is generated. Further, a drive signal for the inverter 3 is generated according to the generated PWM signal.

The inverter 3 includes, for example, two switching elements (for example, power semiconductor elements such as IGBTs and MOS-FETs) for each phase. The supplied direct current is converted into alternating current, and a desired current is passed through the electric motor 4.

The electric motor 4 generates a driving force by the alternating current supplied from the inverter 3, and transmits the driving force to the left and

right driving wheels

9 a and 9 b via the speed reducer 5 and the driving shaft (drive shaft) 8. Further, when the vehicle is driven and rotated by the

drive wheels

9a and 9b, the kinetic energy of the vehicle is recovered as electric energy by generating a regenerative driving force. In this case, the inverter 3 converts an alternating current generated during the regenerative operation of the electric motor 4 into a direct current and supplies the direct current to the battery 1.

The current sensor 7 detects the three-phase alternating currents iu, iv, iw flowing through the electric motor 4. However, since the sum of the three-phase alternating currents iu, iv, and iw is 0, any two-phase current may be detected, and the remaining one-phase current may be obtained by calculation.

The rotation sensor 6 is, for example, a resolver or an encoder, and detects the rotor phase α of the electric motor 4.

FIG. 2 is a flowchart showing a process flow of motor current control performed by the electric motor controller 2.

In step S201, a signal indicating the vehicle state is input. Here, the vehicle speed V (km / h), the accelerator opening AP (%), the rotor phase α (rad) of the electric motor 4, the rotational speed Nm (rpm) of the electric motor 4, and the three-phase AC flowing through the electric motor 4 Currents iu, iv, iw, and a DC voltage value Vdc (V) between the battery 1 and the inverter 3 are input.

The vehicle speed V (km / h) is acquired by communication from another controller such as a vehicle speed sensor (not shown) or a brake controller (not shown). Alternatively, the vehicle rotational speed ωm is multiplied by the tire dynamic radius R, and the vehicle speed v (m / s) is obtained by dividing by the gear ratio of the final gear, and unit conversion is performed by multiplying by 3600/1000 to obtain the vehicle speed V (Km / h) is obtained.

Accelerator opening AP (%) is acquired from an accelerator opening sensor (not shown), or is acquired by communication from another controller such as a vehicle controller (not shown).

The rotor phase α (rad) of the electric motor 4 is acquired from the rotation sensor 6. The rotational speed Nm (rpm) of the electric motor 4 is obtained by dividing the angular speed ω (electrical angle) of the rotor by the number of pole pairs of the electric motor 4 to obtain a motor rotational speed ωm (rad / s) is obtained by multiplying the obtained motor rotational speed ωm by 60 / (2π). The angular velocity ω of the rotor is obtained by differentiating the rotor phase α.

The currents iu, iv, iw (A) flowing through the electric motor 4 are acquired from the current sensor 7.

DC voltage value Vdc (V) is obtained from a voltage sensor (not shown) provided on a DC power supply line between battery 1 and inverter 3 or a power supply voltage value transmitted from a battery controller (not shown).

In step S202, a first torque command value Tm1 ^* , which is a basic target torque command value, is set. Specifically, the first torque command value Tm1 ^* is set by referring to the accelerator opening-torque table shown in FIG. 3 based on the accelerator opening AP and the vehicle speed V input in step S201.

In step S203, the vehicle longitudinal vibration suppression filter process for suppressing the natural vibration frequency component in the unsprung front-rear direction is performed on the first torque command value Tm1 ^* set in step S202, whereby the second torque command value Tm1 ^{* is} set. Torque command value Tm2 ^* is calculated. The unsprung member includes a motor unit including a suspension member. Details of the vehicle longitudinal vibration suppression filter processing will be described later.

In step S204, based on the second torque command value Tm2 ^* set in step S203 and the motor rotation speed ωm, the vibration of the driving force transmission system (the driving shaft 8 A third torque command value Tm3 ^* that suppresses (torsional vibration) is calculated. A detailed calculation method of the third torque command value Tm3 ^* will be described later.

In step S205, the d-axis current target value id ^* and the q-axis current target value iq ^* are obtained based on the third torque command value Tm3 ^* , the motor rotation speed ωm, and the DC voltage value Vdc calculated in step S204.

In step S206, current control is performed to match the d-axis current id and the q-axis current iq with the d-axis current target value id ^* and the q-axis current target value iq ^* obtained in step S205, respectively. For this reason, first, the d-axis current id and the q-axis current iq are obtained based on the three-phase AC current values iu, iv, iw input in step S201 and the rotor phase α of the electric motor 4. Subsequently, d-axis and q-axis voltage command values vd and vq are calculated from a deviation between the d-axis and q-axis current command values id ^* and iq ^* and the d-axis and q-axis current id and iq.

Next, three-phase AC voltage command values vu, vv, and vw are obtained from the d-axis and q-axis voltage command values vd and vq and the rotor phase α of the electric motor 4. Then, PWM signals tu (%), tv (%), and tw (%) are obtained from the obtained three-phase AC voltage command values vu, vv, and vw and the DC voltage value Vdc. The electric motor 4 can be driven with a desired torque indicated by the torque command value by opening and closing the switching element of the inverter 3 by the PWM signals tu, tv, and tw obtained in this way.

Details of the vehicle longitudinal vibration suppression filter processing performed in step S203 of FIG. 2 will be described.

FIG. 4A and FIG. 4B are views showing an unsprung structure on the lower side of the suspension spring, FIG. 4A is a perspective view, and FIG. 4B is a top view. As shown in FIG. 4A, the electric motor 5 is provided under the spring.

The unsprung longitudinal vibration is excited when the reaction force of the driving force transmitted from the road surface vibrates the motor unit including the suspension member back and forth via the lower arm (see FIG. 4B). Therefore, the acceleration sensed by the passenger can be approximated by the sum of the following (A) and (B).
(A) Acceleration generated when the vehicle body is propelled by the driving force (B) Acceleration disturbance generated when the unsprung portion is vibrated back and forth by the driving force

Here, the acceleration disturbance (B) can be expressed by the following equation (1) by parameter identification or the like.

Gv (s) in the formula (1) is represented by the following formula (2). However, F in Formula (1) is a driving force, αv is an acceleration disturbance, and ζv and ωv in Formula (2) are a damping coefficient and natural vibration frequency of unsprung longitudinal vibration, respectively.

In the following, a method for deriving a vehicle longitudinal vibration suppression filter for deriving a transfer characteristic from the basic target torque command value to the acceleration disturbance represented by Equation (1) and canceling the unsprung longitudinal vibration component present in the transfer characteristic is described. Will be described.

If the dead zone due to gear backlash is expressed by the difference between the linear function and the saturation function, the equation of motion of the vehicle is expressed by the following equations (3) to (8).

Here, each parameter in the equations (3) to (8) is as follows.
Jm: Motor inertia Jw: Drive wheel inertia (for one wheel)
M: Vehicle weight Kd: Torsional rigidity of drive shaft Kt: Coefficient related to friction between tire and road surface Nal: Overall gear ratio r: Tire load radius ωm: Motor rotation speed ωw: Angular speed Tm of drive wheel: Motor torque Td: Drive Shaft shaft torque F: Vehicle driving force V: Vehicle speed θ: Torsion angle of drive shaft

However, St ((theta)) in Formula (6) is a saturation function, and is represented by following Formula (9). In formula (9), θBL is the amount of gear backlash in the overall from the motor 4 to the drive shaft (drive shaft) 8.

When equations (3) to (8) are Laplace transformed, they are represented by the following equations (10) to (15).

From the equations (10) to (15), the transfer characteristic from the motor torque Tm to the motor rotation speed ωm is expressed by the following equation (16). However, Gp (s) and Gps (s) in Expression (16) are represented by Expressions (17) and (18), respectively.

However, each parameter in the equations (17) and (18) is represented by the following equation (19).

From equations (10) to (15), the transmission characteristic from the motor torque Tm to the twist angle θ of the drive shaft is expressed by the following equation (20). However, Gt (s) and Fs (s) in Formula (20) are represented by Formula (21) and Formula (22), respectively.

However, p1 and p0 in Formula (21) and Formula (22) are represented by the following Formula (23). Also, ζp and ωp are the damping coefficient and natural vibration frequency of the drive torque transmission system, respectively.

From the equations (13) and (20), the drive shaft shaft torque Td is expressed by the following equation (24).

Here, when the vibration suppression control performed in step S204 of FIG. 2 is applied, the drive shaft shaft torque is expressed by the following equations (25) and (26).

Here, the torque command value Tm ^** in the equation (25) is a second torque command value Tm2 ^* that is a torque command value after the vehicle longitudinal vibration suppression filter is applied (in step 204 of FIG. 2, vibration suppression control is performed). Is the torque command value).

Next, from the equations (11), (12), and (14), the transmission characteristics from the drive shaft shaft torque Td to the driving force F are expressed by the following equations (27) and (28).

However, kf in Formula (28) is represented by following Formula (29).

Furthermore, since the transmission characteristic from the driving force F to the acceleration disturbance αv is expressed by the equations (1) and (2), the torque command value before damping control is obtained from the equations (1), (25), and (27). The transfer function from Tm ^** to the unsprung longitudinal acceleration αv is expressed by the following equation (30).

Here, the reference response αvm of the unsprung longitudinal acceleration expressed by the equation (30) is expressed by the following equation (31). Gvm (s) in the formula (31) is represented by the following formula (32). Tm ^* is a basic torque command value (first torque command value Tm1 ^* ).

Assuming that αv shown in the equation (30) and αvm shown in the equation (31) are αv = αvm, the torque command value Tm ^** is expressed by the following equation (33).

However, Hv (s) in Formula (33) is represented by the following Formula (34).

In general vehicle characteristics, an approximate expression of the following expression (35) is established, and therefore, expression (34) is rewritten to the following expression (36) by pole-zero cancellation.

However, since the expression (36) is non-proper, by adding a low-pass filter with a time constant τ, the expression (36) can be expressed by the third order / third order = (first order / Expressed by a filter of (primary) x (secondary x secondary)

時 The time constant τ should be a sufficiently small value and set to an appropriate value by simulation or experiment.

The drive shaft twist angle θ in the equation (33) may be calculated from the equation (20) or may be obtained from the following equation (38). Equation (38) is obtained by obtaining a torque command value such that the drive shaft torque Td and the motor torque Tm coincide with each other and substituting the obtained torque command value into Equation (20).

FIG. 5 is a diagram illustrating a configuration of the vehicle longitudinal vibration suppression filter. The torque command value Tm ^** (second torque command value Tm2 ^* ) is expressed by the equation (33). FIG. 5 shows the second torque command value by inputting the first torque command value Tm1 ^*. The structure of the vehicle longitudinal vibration suppression filter which outputs Tm2 ^* is shown. As shown in FIG. 5, the vehicle longitudinal vibration suppression filter is represented by a control block 501 represented by (s ² + 2ζv · ωv · s + ωv ² ) / (s ² + 2ωv · s + ωv ² ) and Expression (9). A control block 503 having a saturation function (limiter) 502, a transfer characteristic Hv (s) represented by Expression (37), and a control block 504 having a transfer characteristic Fs (s) represented by Expression (22). With.

Next, the process performed in step S204 in FIG. 2, that is, a method for calculating the third torque command value Tm3 ^* will be described.

FIG. 6 is an example of a control block diagram for performing processing for calculating the third torque command value Tm3 ^* . The vibration suppression control calculation unit 600 that calculates the third torque command value Tm3 ^* includes a vibration suppression control FB calculation unit 600A, a vibration suppression control FF calculation unit 600B, and an adder 600C.

First, the configuration of the vibration suppression control FF calculation unit 600B will be described.

When a torque command value is obtained such that the drive shaft shaft torque Td and the motor torque Tm coincide with each other, the following equations (39) and (40) are obtained.

Therefore, a linear filter G _INV (s) that reduces the natural vibration frequency component of the torque transmission of the vehicle, a filter Gt (s) that calculates the drive shaft torsion angle, a saturation function (limiter), and a wheel inertia of the drive shaft torsion angle The configuration of the vibration suppression control FF calculation unit 600B is represented in FIG. 7 by the filter Fs (s) that compensates for the phase shift due to the tire friction force. That is, the vibration suppression control FF calculation unit 600B includes a control block 601 having a transfer characteristic G _INV (s), a control block 602 having a transfer characteristic Gt (s), and a characteristic represented by the equation (9). A limiter 603, a control block 604 having a transfer characteristic of Fs (s), an adder 605, an adder 606, and a subtractor 607 are provided.

By substituting equation (39) into equation (20), equivalent conversion to the following equation (41) can be made.

Accordingly, the vibration suppression control FF calculation unit 600B includes a linear filter G _INV (s) for reducing the natural vibration frequency component of the torque transmission of the vehicle, a filter Gtm (s) for calculating the ideal response of the drive shaft torsion angle, and saturation. A function (limiter), a wheel inertia of a drive shaft torsion angle, and a filter Fs (s) that compensates for a phase shift due to tire frictional force can also be configured as shown in FIG. According to the configuration shown in FIG. 8, the vibration suppression control FF calculation unit 600B includes a control block 601 having a transfer characteristic G _INV (s), a limiter 603 having a characteristic represented by Expression (9), and Fs (s ), A control block 604 having a transfer characteristic, an adder 606, a subtractor 607, a control block 612 having a transfer characteristic Gtm (s), and an adder 611.

Subsequently, feedback calculation processing performed by the vibration suppression control FB calculation unit 600A will be described. FIG. 9 is a detailed block diagram of the vibration suppression control FB calculation unit 600A. The vibration suppression control FB operation unit 600A has a control block 651 having a transfer characteristic Gp (s), a control block 652 having a transfer characteristic Gps (s), and a transfer characteristic H (s) / Gp (s). A control block 653, an adder 654, a subtractor 655, a control block 656 having a transfer characteristic Gt (s), a limiter 657 having a characteristic expressed by the equation (9), and a transfer Fs (s). A control block 658 having characteristics and an adder 659 are provided.

Gp (s) is a linear plant model indicating the transfer characteristic of the motor rotation speed with respect to the motor torque input to the vehicle, and Gps (s) is a transfer function for calculating the backlash compensation amount of the motor rotation speed. Gt (s) is a filter that calculates the drive shaft twist angle, and Fs (s) is a filter that compensates for a phase shift caused by wheel inertia and tire friction force of the drive shaft twist angle. ), (22).

The vibration suppression control FB calculation unit 600A receives the post-vibration control torque command value Tm3 ^* and the output FF _out of the vibration suppression control FF calculation unit 600B, and receives the linear plant model Gp (s), the filter Gt (s), and the filter Fs. (S), a saturation function (limiter), and a transfer function Gps (s), a motor rotational speed estimated value ωm ^ is calculated. Further, the difference between the calculated motor rotation speed estimated value ωm ^ and the actual motor rotation speed ωm is input, and the output FB _out of the vibration suppression control FB calculation unit 600A is calculated from the transfer function H (s) / Gp (s). .

The transfer function H (s) is a feedback element that reduces only vibration when a band-pass filter is used. At this time, the greatest effect can be obtained by setting the filter characteristics as shown in FIG. That is, the transfer function H (s) has substantially the same attenuation characteristics on the low-pass side and the high-pass side, and the torsional resonance frequency of the drive system is close to the center of the passband on the logarithmic axis (log scale). Is set to For example, when H (s) is composed of a first-order high-pass filter and a first-order low-pass filter, the frequency fp is the torsional resonance frequency of the drive system, k is an arbitrary value, and the equation (42) Constitute.

_{However, τ L = 1 / (2πf} HC), f HC = k · f p, τ H = 1 / (2πf LC), it is f _LC = f _p / k.

Hereinafter, the linear plant model Gp (s) indicating the transfer characteristic of the motor rotation speed with respect to the motor torque input to the vehicle and the transfer function Gps (s) for calculating the backlash compensation amount of the motor rotation speed will be described.

As described above, the transfer characteristics from the motor torque Tm to the motor rotational speed ωm are expressed by the equations (16) to (18).

Formula (17) is arranged and expressed as the following formula (43). In a general vehicle, when the poles and zeros of the transfer function of Equation (43) are examined, one pole and one zero show extremely close values. This corresponds to the fact that α and β in Equation (43) show extremely close values. Here, ζp and ωp are the damping coefficient and natural vibration frequency of the drive torsional vibration system, respectively.

Therefore, by performing pole-zero cancellation (approximate with α = β) in equation (43), the (second order) / (third order) transfer characteristic Gp (s) is configured as shown in the following equation (44). To do.

As described above, according to the control apparatus for an electric vehicle in the first embodiment, in the control apparatus for an electric vehicle that sets the motor torque command value based on the vehicle information and controls the torque of the motor connected to the drive wheels, A filtering process for reducing the natural vibration frequency component in the unsprung direction in the unsprung direction is performed on the value, and the motor torque is controlled according to the final torque command value obtained by performing the filtering process on the motor torque command value. Thereby, vibration in the front-rear direction under the vehicle spring can be reduced, and smooth acceleration can be realized without causing the occupant to feel a shock or unpleasant vibration.

Further, by using a filter (control block 501 in FIG. 5) including a reverse characteristic of the unsprung front-rear direction transfer characteristic as a filter used in the filtering process for reducing the unsprung front-rear natural vibration frequency component. It is possible to accurately calculate a compensation torque that reduces vibration in the front-rear direction under the vehicle spring. As a result, smooth acceleration can be realized without causing the occupant to feel a shock or unpleasant vibration even in a scene of start acceleration accompanied by a steep torque change.

In particular, the filter used in the filtering process for reducing the unsprung front-rear natural vibration frequency component includes a filter (control block 501 in FIG. 5) including a reverse characteristic of the unsprung front-rear transmission characteristic, and a drive shaft twist. Saturation function 502 with angle as input, filter that outputs correction torque in dead zone due to gear (control block 503), filter that calculates phase shift due to wheel inertia and tire friction force of drive shaft twist angle (control block 504) Including. As a result, even during acceleration from coasting or deceleration, the natural vibration frequency component in the unsprung front-rear direction of the driving force transmission system of the vehicle having the dead zone due to the gear can be suppressed, and vibration in the unsprung front-rear direction can be suppressed. In addition, since it is not necessary to change the filter configuration depending on the presence or absence of the dead band due to the saturation function (limiter), it is not necessary to perform complicated calculations (initialization, condition determination, switching, etc.).

Since the upper and lower limit values of the saturation function 502 are determined based on the gear backlash amount (see equation (9)), the backlash correction amount can be calculated.

Further, a filtering process for suppressing the torsional vibration of the drive shaft (FIG. 2) with respect to the motor torque command value subjected to the filtering process (S203 in FIG. 2) for reducing the natural vibration frequency component in the front-rear direction under the spring. The motor torque is controlled using the motor torque command value subjected to the filtering process (vibration control) for suppressing the torsional vibration of the drive shaft as the final torque command value. The damping control of the driving force transmission system may be configured in consideration of a detection delay of the motor rotation speed, assuming a model to be controlled from the motor torque to the motor rotation speed. In such a case, if the filtering process to reduce the natural vibration frequency component in the front-rear direction under the spring is arranged after the filtering process (vibration control) to suppress the torsional vibration of the drive shaft, it is considered in the vibration suppression control. This results in a response different from that of the model to be controlled, and there is a difference between the expected response of the motor rotational speed and the detection delay. As a result, the performance of the vibration suppression control may be deteriorated. On the other hand, as in this embodiment, the filtering process for reducing the unsprung natural vibration frequency component in the front-rear direction is performed in the previous stage of the vibration suppression control, thereby achieving the effect of the vibration suppression control of the driving force transmission system vibration. The vibration in the front-rear direction under the spring of about 20 to 40 Hz can be suppressed.

-Second Embodiment-
The unsprung longitudinal vibration is excited when the reaction force of the driving force transmitted from the road surface is vibrated back and forth through the lower arm, so the acceleration sensed by the occupant is shown in FIG. Can be modeled with a spring, mass, damper system.

The equation of motion of the unsprung spring-mass-damper model configuration diagram shown in FIG. 11 is expressed by the following equation (45).

From Expression (45), the transfer characteristic from the driving force F to the displacement amount xv of the vehicle (vehicle compartment) is expressed by the following Expression (46).

Therefore, the transfer characteristic from the driving force F to the vehicle longitudinal acceleration αv is expressed by the following equation (47). Further, Gv (s) in the formula (47) is represented by the formula (48).

However, v1, v0, ωv, and ζv in Expression (48) are expressed by Expression (49).

Hereinafter, from the first torque command value Tm1 ^* , which is the basic target torque command value, the transfer characteristic up to the vehicle longitudinal acceleration αv represented by the equation (47) is derived, and the unsprung longitudinal vibration existing in the transfer characteristic. A method for deriving a vehicle longitudinal vibration suppression filter that cancels components will be described.

If the dead zone due to gear backlash is expressed by the difference between the linear function and the saturation function, the equation of motion of the vehicle is expressed by the following equations (50) to (55).

Here, each parameter in the equations (50) to (55) is as follows.
Jm: Motor inertia Jw: Drive wheel inertia (for one wheel)
M: Vehicle weight Kd: Torsional rigidity of drive shaft Kt: Coefficient related to friction between tire and road surface Nal: Overall gear ratio r: Tire load radius ωm: Motor rotation speed ωw: Angular speed Tm of drive wheel: Motor torque Td: Drive Wheel torque F: Vehicle driving force V: Vehicle speed θ: Torsion angle of drive shaft

However, St ((theta)) in Formula (53) is a saturation function, and is represented by following Formula (56). In the equation (56), θBL is a gear backlash amount in the overall from the motor 4 to the drive shaft (drive shaft) 8.

When Laplace transform is performed on the equations (50) to (55), the following equations (57) to (62) are obtained.

From the equations (57) to (62), the transfer characteristic from the motor torque Tm to the motor rotational speed ωm is expressed by the following equation (63). However, Gp (s) and Gps (s) in Expression (63) are expressed by Expressions (64) and (65), respectively.

However, each parameter in the equations (64) and (65) is represented by the following equation (66).

From equations (57) to (62), the transmission characteristic from the motor torque Tm to the twist angle θ of the drive shaft is expressed by the following equation (67). However, Gt (s) and Fs (s) in Formula (67) are represented by Formula (68) and Formula (69), respectively.

However, p1 and p0 in Formula (68) and Formula (69) are represented by the following Formula (70). Also, ζp and ωp are the damping coefficient and natural vibration frequency of the drive torque transmission system, respectively.

From the equations (60) and (67), the drive shaft shaft torque Td is expressed by the following equation (71).

Here, when the vibration suppression control performed in step S204 of FIG. 2 is applied, the drive shaft shaft torque is defined by the following equations (72) and (73).

Here, the torque command value Tm ^** in the equation (72) is a second torque command value Tm2 ^* that is a torque command value after the vehicle longitudinal vibration suppression filter is applied (in step 204 of FIG. 2, vibration suppression control is performed). Is the torque command value).

Next, from the equations (58), (59), and (61), the transmission characteristics from the drive shaft shaft torque Td to the driving force F are expressed by the following equations (74) and (75).

However, kf in Formula (75) is represented by following Formula (76).

Further, since the transmission characteristic from the driving force F to the vehicle longitudinal acceleration αv is expressed by the equations (47) and (48), the torque command before vibration suppression control is obtained from the equations (47), (72), and (74). The transfer function from the value Tm ^** to the vehicle longitudinal acceleration αv is expressed by the following equation (77).

Here, the reference response αvm of the vehicle longitudinal acceleration expressed by the equation (77) is expressed by the following equation (78). Gvm (s) in the formula (78) is represented by the following formula (79). Tm ^* is a basic torque command value.

Assuming that αv shown in the equation (77) and αvm shown in the equation (78) are αv = αvm, the torque command value Tm ^** is expressed by the following equation (80).

However, Hv (s) in Formula (80) is represented by the following Formula (81).

In general vehicle characteristics, the approximate expression of the following expression (82) is established. Therefore, the expression (81) is rewritten to the following expression (83) by pole-zero cancellation.

However, since the equation (83) is non-proper, by adding a low-pass filter with a time constant τ, the equation (83) can be expressed as the third order / third order = (first order / It is represented by a filter of (primary) × (secondary × secondary).

The drive shaft twist angle θ in the equation (80) may be calculated from the equation (67) or may be obtained from the following equation (85). Expression (85) is obtained by obtaining a torque command value such that the drive shaft torque Td and the motor torque Tm coincide with each other and substituting the obtained torque command value into Expression (67).

The configuration of the vehicle longitudinal vibration suppression filter in the control apparatus for an electric vehicle in the second embodiment is shown in FIG. 5 as in the first embodiment.

As described above, in the control apparatus for the electric vehicle in the second embodiment, as in the control apparatus for the electric vehicle in the first embodiment, the vibration in the front-rear direction under the vehicle spring is reduced, and the shock and unpleasant vibration are given to the occupant. Smooth acceleration can be realized without feeling.

The control result by the control device for the electric vehicle in the first and second embodiments will be described. First, the acceleration response when the torque command value is increased stepwise from the creep state (without gear backlash) during start-up and rapid acceleration will be described with reference to FIGS. 12A and 12B.

FIG. 12A is a diagram showing an example of a control result by the control device for the electric vehicle in the first and second embodiments. FIG. 12A shows, in order from the top, the time change of the torque command value, the time change of the vehicle longitudinal acceleration, the time change of the drive shaft shaft torque, and the time change of the motor rotation speed after the vibration damping control is performed.

When the basic target torque command value suddenly rises by stepping on the accelerator step by step at time t1 from the stopped state, the drive shaft torque converges to a substantially steady value at time t2. As for the acceleration, the vibration of about 20 to 40 Hz, which is seen in the conventional example described later, is suppressed, rises smoothly, and converges to a substantially steady value at time t2.

FIG. 12B is a diagram illustrating an example of a control result by the control device described in Japanese Patent Laid-Open No. 2003-9566. When the basic target torque command value suddenly rises by stepping on the accelerator at time t1 from the creep state, the torque command value after vibration suppression control is obtained by vibration suppression control to suppress vibration of the driving force transmission system. The torque command value is obtained by suppressing the vibration frequency component of the driving force transmission system with respect to the basic target torque command value, and the drive shaft torque rises without stepwise vibration and converges without reaching time t2. On the other hand, in the acceleration, vibrations of about 20 to 40 Hz remain from the rising time t1 to the time t2.

Next, the response of acceleration when the torque command value is increased stepwise during sudden acceleration from the coast state (with gear backlash) will be described with reference to FIGS. 13A and 13B.

FIG. 13A is a diagram illustrating an example of a control result by the control device for the electric vehicle in the first and second embodiments. FIG. 13A shows, in order from the top, the time change of the torque command value, the time change of the vehicle longitudinal acceleration, the time change of the drive shaft shaft torque, and the time change of the motor rotation speed after the vibration damping control is performed.

When the basic target torque command value suddenly rises by stepping on the accelerator step by step at time t1 from the stopped state, the drive shaft torque converges to a substantially steady value at time t2. As for the acceleration, the vibration of about 20 to 40 Hz, which is seen in the conventional example described later, is suppressed, rises smoothly, and converges to a substantially steady value at time t2. That is, vibration in the front-rear direction under the spring can be suppressed even in a scene with gear backlash.

FIG. 13B is a diagram illustrating an example of a control result by the control device described in Japanese Patent Laid-Open No. 2003-9566. When the basic target torque command value suddenly rises by stepping on the accelerator at time t1 from the coast state, in order to correct the error of the dead zone of the gear backlash, by generating the torque command value on the acceleration side, From time t1 to time t2, vibration (shock) of acceleration of about 20 to 40 Hz is promoted.

Thus, according to the control device for the electric vehicle in the first and second embodiments, regardless of the presence or absence of the backlash of the gear, the occupant feels shock and unpleasant vibration, while being smooth, A steep acceleration performance without impairing the response can be realized.

The present invention is not limited to the first and second embodiments described above, and various modifications and applications can be made without departing from the gist of the present invention. For example, the drive shaft torsion angle θ used in the filtering process for reducing the unsprung front-rear natural vibration frequency component suppresses the vibration of the driving force transmission system (torsional vibration of the drive shaft 8) performed in step S204 of FIG. You may use the value calculated by the filtering process to do.

This application claims priority based on Japanese Patent Application No. 2012-222371 filed with the Japan Patent Office on October 4, 2012, the entire contents of which are incorporated herein by reference.

Claims

In a control device for an electric vehicle that sets a motor torque command value based on vehicle information and controls the torque of a motor connected to a drive wheel,
Filtering means for applying a filtering process to reduce the natural vibration frequency component of the unsprung front and rear direction with respect to the motor torque command value;
Motor torque control means for controlling the motor torque in accordance with a final torque command value obtained by applying the filtering process to the motor torque command value;
The control apparatus of the electric vehicle provided with.
In the control apparatus of the electric vehicle according to claim 1,
The filter used in the filtering process is a filter including a reverse characteristic of the unsprung front-rear transmission characteristic.
Control device for electric vehicle.
In the control apparatus of the electric vehicle according to claim 2,
The filter used in the filtering process includes a filter including a reverse characteristic of the unsprung front-rear transmission characteristic, a saturation function that receives a drive shaft torsion angle, a filter that outputs a correction torque in a dead zone by a gear, and a drive Including a wheel inertia of a shaft twist angle and a filter for calculating a phase shift due to tire friction force,
Control device for electric vehicle.
In the control apparatus of the electric vehicle according to claim 3,
The upper and lower limit values of the saturation function are determined based on the gear backlash amount.
Control device for electric vehicle.
In the control apparatus of the electric vehicle as described in any one of Claims 1-4,
The motor torque command value subjected to the filtering process further includes a torsional vibration suppressing means for performing a filtering process for suppressing the torsional vibration of the drive shaft,
The motor torque control means controls the motor torque with a motor torque command value subjected to filtering processing for suppressing torsional vibration of the drive shaft as a final torque command value;
Control device for electric vehicle.
In a control method for an electric vehicle that sets a motor torque command value based on vehicle information and controls the torque of a motor connected to a drive wheel,
Applying a filtering process to reduce the natural vibration frequency component of the unsprung longitudinal direction to the motor torque command value,
Controlling the motor torque according to a final torque command value obtained by performing the filtering process on the motor torque command value;
Control method of electric vehicle.