JP5928046B2 - Wheel type moving body - Google Patents

Description

  The present specification relates to a wheel type moving body that travels on a road surface by driving a wheel.

  Patent Document 1 discloses a wheel-type moving body that stabilizes the posture angle of the vehicle body (that is, the posture angle of the vehicle body with respect to the direction of gravity) by inclining the vehicle body with respect to the road surface in accordance with the traveling state. .

JP 2011-11575 A

  The wheel type moving body described above is provided with a vehicle body tilting device (actuator) for tilting the vehicle body with respect to the road surface. The vehicle body tilting device (actuator) needs to generate a large torque to tilt the vehicle body, and usually includes a high speed reduction mechanism. For this reason, the disturbance force represented by the frictional force is large, and it is difficult to control the vehicle body to a desired posture angle as it is. The present specification discloses a technique for compensating for a disturbance force acting on an actuator for tilting the vehicle body and controlling the vehicle body to a desired posture angle.

  The wheel type moving body disclosed in the present specification includes a vehicle body, means for detecting or estimating a posture angle around the motion axis of the vehicle body, an actuator capable of adjusting the posture angle around the motion axis of the vehicle body, and an output shaft of the actuator. , A first control value obtained by processing the detected or estimated posture angle of the vehicle body with a low-pass filter, and a second control value obtained by processing the detected rotation angle of the output shaft of the actuator with a high-pass filter. And a control device for controlling the actuator.

  The wheel-type moving body uses a first control value obtained by processing the posture angle of the vehicle body with a low-pass filter and a second control value obtained by processing the rotation angle of the output shaft of the actuator with a high-pass filter. As will be described later, if the posture angle of the vehicle body is used, it can be suitably compensated for disturbance forces in the low frequency range, but it should be suitably compensated for disturbance forces in the high frequency range by estimating the posture angle. Is difficult. On the other hand, when the rotation angle of the output shaft of the actuator is used, it is possible to suitably compensate for disturbance forces in a high frequency region. In the above-described wheel type moving body, by using the first control value and the second control value, the disturbance force acting on the actuator can be compensated in a wide band (that is, a low frequency region and a high frequency region), and the vehicle body can have a desired posture angle. Can be suitably controlled

The front view which shows typically the structure of the vehicle which concerns on an Example. The side view of the vehicle shown in FIG. The block diagram which shows the structure of the control system which concerns on an Example. The figure which modeled the vehicle which relates to the execution example The block diagram which shows the structure of the disturbance compensation of the control system which concerns on an Example. The block diagram which added the feedback control loop which feeds back the attitude angle of a vehicle body to the control system shown in FIG. The block diagram which shows the structure of the disturbance compensation which concerns on a comparative example. The figure which shows the result of a numerical simulation. The block diagram which shows the structural example of the feedback control system which feeds back a 1st control value and a 2nd control value directly. The block diagram which shows the other structural example of the feedback control system which feeds back a 1st control value and a 2nd control value directly.

  The main features of the embodiments described below are listed. The technical elements described below are independent technical elements and exhibit technical usefulness alone or in various combinations, and are not limited to the combinations described in the claims at the time of filing. Absent.

(Characteristic 1) In the embodiment disclosed in the present specification, the control device feedback-compensates feedback to the actuator so that the posture angle of the vehicle body becomes the target posture angle based on the detected or estimated posture angle of the vehicle body. May have a part. According to such a configuration, the posture angle of the vehicle body can be suitably controlled to the target posture angle.

(Characteristic 2) In the embodiment disclosed in the present specification, the control device determines the difference between the second-order differential value of the sum of the first control value and the second control value and the control input to the actuator as a disturbance generated in the actuator. A disturbance compensation unit that compensates for the above may be included. According to such a configuration, the influence of the disturbance generated in the actuator is compensated, and the posture angle of the vehicle body can be suitably controlled.

(Feature 3) In the embodiment disclosed in this specification, the break frequency of the low-pass filter may be lower than the natural vibration frequency of the vehicle body, and the break frequency of the high pass filter may be lower than the natural vibration frequency of the vehicle body. According to such a configuration, the influence of the resonance of the vehicle body can be reduced, and the posture angle of the vehicle body can be suitably controlled.

  The present embodiment will be described with reference to the drawings. The present embodiment is an example in which the technique disclosed in this specification is applied to the roll angle control around the roll axis of the vehicle body, and the configuration not related to the roll angle control is the same as that of the conventional technique. For this reason, in the following description, the structure relevant to roll angle control is mainly demonstrated, and description is abbreviate | omitted suitably about the other structure. As shown in FIGS. 1 and 2, the vehicle 100 includes a vehicle body 10 to which left and right front wheels 12 a and 12 b, one rear wheel 15 (driven wheel), and front wheels 12 a and 12 b and a rear wheel 16 are attached. .

  The front wheels 12 a and 12 b are rotatably attached to the side surface of the vehicle body 10. The front wheels 12a and 12b are independently driven by in-wheel motors 14a and 14b. The vehicle 10 travels on the traveling surface R by driving the front wheels 12a and 12b. The rear wheel 15 is attached to the vehicle body 10 and is disposed at a position that becomes the center of the vehicle body 10 when the vehicle 100 is viewed from the front. In this embodiment, an omnidirectional wheel or a caster wheel is used for the rear wheel 15. For this reason, by controlling the rotational drive amount (rotational angular velocity) of the left and right front wheels 12a and 12b, the vehicle 100 can change its traveling direction to an arbitrary direction. Note that the rear wheel 15 may be a steering wheel, and the traveling direction of the vehicle 100 may be controlled by the steering angle of the rear wheel 15.

  The vehicle body 10 is rotatable around the roll axis. The vehicle body 10 includes a gyro sensor 16 that detects the roll angular velocity of the vehicle body 10, an actuator 18 that controls the roll angle of the vehicle body 10, and a control device 20 (illustrated in FIG. 3) that controls the actuator 18.

  The gyro sensor 16 detects the roll angular velocity of the vehicle body 10. As shown in FIG. 3, the gyro sensor 16 is electrically connected to the control device 20. The roll angular velocity signal output from the gyro sensor 16 is input to the control device 20.

  The actuator 18 is a device for adjusting the roll angle of the vehicle body 10. The actuator 18 applies a torque τ around the roll axis to the vehicle body 10. As shown in FIG. 3, the actuator 18 is electrically connected to the control device 20. Based on the control command value output from the control device 20, the actuator 18 generates a torque τ around the roll axis. In this embodiment, a geared servo motor is used for the actuator 18. The geared servo motor generates a large torque τ by decelerating and outputting the output from the motor by a reduction mechanism (gear). The rotation angle of the output shaft of the actuator 18 is detected by the encoder 22. The rotation angle of the output shaft of the actuator 18 detected by the encoder 22 is input to the control device 20.

  The control device 20 is configured by a computer including a CPU, a ROM, a RAM, and the like. The control device 20 is installed in the vehicle body 10. The control device 20 is electrically connected to the gyro sensor 16, the actuator 18, and the encoder 22. The control device 20 drives the actuator 14 based on the roll angular velocity input from the gyro sensor 16 and the rotation angle of the output shaft of the actuator 18 input from the encoder 22. Thus, the posture angle (roll angle) around the roll axis of the vehicle body 10 is controlled. Hereinafter, the control device 20 will be described in detail.

First, a motion model around the roll axis of the vehicle body 10 will be described, and then a control configuration of the control device 20 will be described based on the motion model. As shown in FIG. 4, the vehicle body 10 is represented by an arm 10a suspended by suspensions 13a and 13b and a vehicle body upper portion 10b attached to the arm 10a. In this embodiment, the rigidity of the front wheels 12a and 12b is ignored as being sufficiently high, and the arm 10a is supported on the road surface R via the left and right suspensions 13a and 13b. The posture angle of the upper body 10b can be controlled by an actuator 18 attached to the arm 10a. Here, J is the moment of inertia of the vehicle body upper portion 10b around the output shaft _{S r} of the actuator 18, M is the mass of the upper body 10b, l is from the output shaft _{S r} of the actuator 18 to the center of gravity of the vehicle body upper 10b It is a distance, θ is a posture angle (attitude angle with respect to the direction of gravity) of the vehicle body upper part 10 b, θ _n is a rotation angle of the arm, and τ represents a torque of the actuator 18.

In the present embodiment, when the road surface R is uneven or the road surface R is inclined, the output shaft of the actuator 18 can freely rotate. Therefore, theoretically, the disturbance on the road surface R does not affect the posture angle of the upper body 10b. However, the actuator 18 has a reduction mechanism with a high gear ratio, and a disturbance force τ _dis (frictional force or the like) acts on the output shaft of the actuator 18. Therefore, the equation of motion of the vehicle body upper portion 10b is as shown in the following equation (1) in consideration of the disturbance force τ _dis acting on the output shaft of the actuator 18.

  Here, in the present embodiment, control is performed so that the posture angle θ of the vehicle body upper portion 10b becomes 0 ° except when the vehicle 100 performs a turning maneuver (in the case of a straight traveling motion or the like). That is, the torque due to gravity Mglsin θ is zero. When the vehicle 100 performs a turning motion, the inclination angle θ of the vehicle body upper portion 10b is controlled so that the torque due to gravity acting on the vehicle body upper portion 10b and the torque due to centrifugal force acting on the vehicle body upper portion 10b are balanced. For this reason, Mglsin θ can be handled as 0 even during the turning motion. Therefore, the above formula (1) is expressed as the following formula (2).

Note that, when the formula (2) is Laplace transformed, the following formula (3) is obtained. As apparent from the equation (3), in order to set the vehicle body upper portion 10b to the desired posture angle θ, the disturbance force τ _dis (s) is estimated, and the estimated disturbance force τ ′ _dis (s) is used. Need to compensate. That is, estimates the disturbance force τ _{dis (s),} it is necessary to drive the actuator 18 so as to cancel the estimated estimated disturbance force τ _{'dis (s).} In order to set the vehicle body upper portion 10b to a desired posture angle θ, it is necessary to accurately estimate the disturbance force τ _dis (s).

Here, as a method for estimating the disturbance force τ _dis (s), for example, an estimation method using a disturbance observer can be considered. Using the twice differential value s ² θ (s) of the attitude angle θ (s) of the upper body 10b, the control input τ (s), and the moment of inertia J of the upper body 10b, the estimated disturbance force τ ′ _dis (s ) Is obtained by the following equation (4). In Expression (4), Q (s) is a low-pass filter having a steady gain of 1 and an order of 2 or more.

As is clear from the equation (4), the estimated disturbance force τ ′ _dis (s) can obtain a good estimation result for frequencies below the break frequency of the low-pass filter Q (s). That is, the low-pass filter Q (s) has a gain of 1 and a phase of 0 deg with respect to frequencies below the breakpoint frequency. Further, as apparent from the equation (4), the estimated disturbance force τ ′ _dis (s) is not affected by the angle θ _n of the arm 10a. Therefore, even if the arm 10a rotates due to the unevenness or inclination of the road surface R, the posture angle θ of the vehicle body upper portion 10b does not change. From the above, for the disturbance τ (s) having a frequency equal to or lower than the break frequency of the low-pass filter Q (s), the posture of the upper body 10b can be accurately determined by using the estimated disturbance force _τ'dis (s). It can be controlled well.

However, the low-pass filter Q (s) resonates due to (1) an estimated delay of the posture angle θ of the upper body 10b, and (2) spring characteristics of the vehicle 100 (rigidity of the front wheels 12a and 12b, rigidity of the suspensions 13a and 13b, etc.). The breakpoint frequency cannot be set to a sufficiently high frequency due to vibration, (3) backlash of the gear of the actuator 18, and the like. Therefore, it is necessary to accurately estimate the disturbance force τ _dis (s) in a frequency region higher than the break frequency of the low-pass filter Q (s).

As another method for estimating the disturbance force τ _dis (s), it is conceivable to use the angle (θ−θ _n ) of the output shaft of the actuator 18 instead of the attitude angle θ of the vehicle body upper part 10b. Since the angle of the output shaft of the actuator 18 is expressed as θ (s) −θ _n (s), if this is substituted into the right side of the equation (4), the estimated disturbance force τ ′ _disa ( s) can be estimated.

To compensate for the effects of the disturbance force _{tau dis} (s) is by the equation estimated disturbance force is estimated by (5) _{tau 'disa} (s), the disturbance force _{tau dis} located on the right-hand side of formula (3) (s) is cancel. Therefore, the following relationship needs to be established between the disturbance force τ _dis (s) and the estimated disturbance force τ ′ _disa (s).

  When the above equation (6) is substituted into the relational equation of equation (3), the following equation (7) is obtained.

  Next, when the relationship of Expression (4) is substituted into Expression (7), the following Expression (8) is obtained.

Here, assuming that the disturbance force τ _dis (s) and the estimated disturbance force τ ′ _dis (s) are equal in the region where the gain of the low-pass filter Q (s) is 1 and the phase is 0 °, the equation (8) ) Becomes the following formula (9).

As is clear from equation (9), in the frequency region lower than the break frequency of the low-pass filter Q (s), when the arm 10a rotates due to road surface disturbance, the posture angle θ of the vehicle body upper portion 10b also changes accordingly. Is shown. Therefore, when the disturbance force τ _dis (s) is estimated using the angle (θ−θ _n ) of the output shaft of the actuator 18, in the frequency region lower than the break frequency of the low-pass filter Q (s), The posture angle θ of the upper body 10b cannot be controlled to a desired posture angle due to the disturbance force.

FIG. 7 shows a block diagram of a control configuration when the disturbance force τ _dis (s) is estimated using the angle (θ−θ _n ) of the output shaft of the actuator 18. As is apparent from FIG. 7, the control input τ and the disturbance force τ _dis are input to the vehicle body upper portion 10b that is the control target, and the rotation angle (θ−θ _n ) of the output shaft of the actuator 18 is fed back. The difference between the fed back value and the control input τ is given a low-pass filter Q (s), and the control input τ is compensated by the value.

Therefore, in the control device 20 of the present embodiment, the attitude angle θ of the vehicle body upper portion 10b is used at a frequency lower than the set frequency f = 2πω, and the rotation angle of the output shaft of the actuator 18 at a frequency higher than the set frequency f = 2πω ( The disturbance force τ _dis is estimated using θ−θ _n ). Therefore, the estimated disturbance force τ ′ _disb (s) estimated by the control device 20 of the present embodiment is expressed by the following equation (10) by _modifying the equation (4).

As is clear from the equation (10), a low-pass filter (ω / s + ω) is given to the posture angle θ (s) of the vehicle body upper portion 10b, and the rotation angle (θ (s) −θ _n ) of the output shaft of the actuator 18 is obtained. (S)) is given a high-pass filter (s / s + ω), and the estimated disturbance force τ ′ _disb (s) is calculated using the sum of these. When the above equation (10) is processed in the same manner as the transformation from the equation (6) to the equation (9), the posture angle θ (s) of the vehicle body upper part 10b is expressed by the following equation (11).

As is clear from the comparison between Expression (11) and Expression (9), in the control device 20 of the present embodiment, the frequency component of θn lower than the set frequency f = 2πω and the break frequency set by Q (s). The above-described road surface disturbance component does not affect the posture angle θ of the vehicle body upper portion 10b. FIG. 5 shows a block diagram of a configuration for compensating the disturbance force τ _dis of the control device 20 of the present embodiment.

In the present embodiment, a control loop that feeds back the attitude angle θ of the vehicle body upper part 10b is provided outside the control loop that compensates for the disturbance force _τdis . That is, the control device 20 of the present embodiment has the configuration shown in FIG. In FIG. 6, “FB” is a feedback compensator and can be designed using PID control, H∞ control, or the like. In the configuration shown in FIG. 6, the order of the low-pass filter (ω / s + ω) and the high-pass filter (s / s + ω) is the first order. However, the order of the low-pass filter and the high-pass filter may be other than the first order. Good.

Next, the operation of the control device 20 described above will be described. Here, the inertia moment J of the vehicle body upper part 10b is 10.0 kgm ² , a step disturbance force τ _dis of 5.0 Nm is applied at the time t = 1.0 s, and the frequency of 0. 0 s after the time t = 10.0 s. The posture angle θ of the vehicle body 10b was simulated under the condition that a road disturbance force of 05 Hz acts. In the simulation, the low-pass filter Q (s) was a secondary low-pass filter having a break frequency at 5.0 Hz, and ω was 2π1.0. As a comparative example, the same simulation was performed for the control configuration shown in FIG. 7 (that is, when only the rotation angle (θ−θ _n ) of the output shaft of the actuator 18 is used).

  FIG. 8 shows the simulation result. The upper part of FIG. 8 shows the step disturbance acting at time t = 1.0 s, the middle part shows the road surface disturbance acting at time t = 10.0 s, and the lower part represents the attitude angle θ of the upper body 10b. As is apparent from FIG. 8, in the comparative example, the step disturbance at time t = 1.0 s is preferably compensated, but the posture angle θ is affected by the road surface disturbance after time t = 10.0 s. It has fluctuated. On the other hand, in the control device 20 of the present embodiment, the posture angle θ is controlled to a constant value without being affected by the step disturbance at the time t = 1.0 s and the road surface disturbance after the time t = 10.0 s. .

As is clear from the above description, in the vehicle 100 of the present embodiment, the disturbance force τ _dis acting on the vehicle body 10 can be accurately estimated over a wide band, and the estimated disturbance force τ ′ _dis can be used as an actuator. 18 is driven. As a result, the posture angle θ of the vehicle body 10 can be stably controlled to a desired posture angle.

  The control technique of the present embodiment can be suitably applied to a small moving body having a short wheel base and a short tread width. That is, in this type of small mobile body, it is required to change the inclination angle of the vehicle body according to the traveling state and to stabilize the posture of the vehicle body. When the control technique of this embodiment is used, the influence of disturbance force can be suppressed and the posture angle of the vehicle body can be stably controlled to a desired angle. As a result, the running stability of the small mobile body can be improved.

  The disturbance compensation method described above is merely an example, and the disturbance compensation method is not limited to the above-described method. Therefore, disturbance compensation may be performed by a configuration obtained by equivalently converting the configuration of FIG.

Finally, the correspondence between the above-described embodiments and the claims will be described. The roll axis of the vehicle body 10 is an example of “vehicle motion axis” in the claims, the gyro sensor 16 is an example of “means for detecting or estimating the attitude angle of the vehicle body”, and the encoder 22 is “an output shaft of the actuator”. [Ω / (s + ω)] · θ (s) is an example of “first control value”, and [s / (s + ω)] · (θ (s) −θ _n (s)) is an example of “second control value”, the configuration shown in FIG. 5 is an example of “disturbance compensation unit”, and FB shown in FIG. 6 is an example of “feedback compensation unit”. .

  Specific examples of the present invention have been described in detail above, but these are merely examples and do not limit the scope of the claims. The technology described in the claims includes various modifications and changes of the specific examples illustrated above.

  For example, the embodiment described above is an example in which the technology disclosed in the present specification is applied to the roll angle control of the vehicle body, but the technology disclosed in the present specification can be applied to other than the roll angle control of the vehicle body, For example, the present invention can be applied to vehicle body pitch angle control. When the technique disclosed in this specification is applied to vehicle body pitch angle control, an actuator for controlling the vehicle body pitch angle and a sensor for detecting the vehicle body pitch angle may be provided. By providing such an actuator and a pitch angle sensor, the technology disclosed in this specification can be applied to the pitch angle control of the vehicle body.

In the above-described embodiment, the first control value (the value obtained by processing the attitude angle θ of the vehicle body with the low-pass filter) and the second control value (the rotation angle (θ−θ _n ) of the output shaft of the actuator) are obtained with the high-pass filter. The disturbance force is estimated using the processed value), but the technique disclosed in the present specification is not limited to such an example. For example, the posture angle θ of the vehicle body may be controlled by directly using these control values for feedback control.

When the first control value and the second control value are directly used for feedback control, for example, the configuration shown in FIG. 9 can be adopted. In the feedback control system shown in FIG. 9, the difference (θ−θ _ref ) between the vehicle body posture angle θ and the target posture angle θ _ref is input to the first feedback compensator FB ₁ . The first feedback compensator FB ₁ calculates an output so that the difference (θ _ref −θ) between the posture angle θ of the vehicle body and the target posture angle θ _ref becomes zero. Also, (1) the output from the first feedback compensator FB ₁ , (2) the first control value [ωθ / (s + ω)] and the second control value [s (θ−θ _n ) / (s + ω)] differential value of the sum (the pseudo posture angular velocity), the difference is inputted to the second feedback compensator FB _2. The second feedback compensator FB ₂ calculates the control input τ to the actuator so that the difference between (1) and (2) becomes zero. The first feedback compensator FB ₁ is a feedback compensator for the posture angle θ of the vehicle body, and can be designed by PID control theory, phase lead / lag compensation, H∞ compensation, and the like. The second feedback compensator FB ₂ is a feedback compensator for the pseudo posture angular velocity of the vehicle body, and can be designed by PID control theory, phase lead / lag compensation, H∞ compensation, and the like. For example, the first feedback compensator FB ₁ can be a P compensator and the second feedback compensator FB ₂ can be a PI compensator. By adopting the configuration shown in FIG. 9, the gain of the second feedback compensator FB ₂ can be set larger than when the rotation angle (θ−θ _n ) of the output shaft of the actuator is not used, The disturbance τ _dis can be suppressed with high accuracy. In the configuration shown in FIG. 9, an observer for estimating the disturbance τ _{dis according} to the embodiment may be further incorporated inside the control loop that feeds back the pseudo posture angular velocity.

Furthermore, when using a 1st control value and a 2nd control value for direct feedback control, the structure shown in FIG. 10 can be taken, for example. In the feedback control system shown in FIG. 10, (a) target attitude angle θ _ref , (b) first control value [ωθ / (s + ω)] and second control value [s (θ−θ _n ) / (s + ω) ] (Pseudo attitude angle) and the difference between the sum and the sum of the values is input to the feedback compensator FB. The feedback compensator FB calculates the control input τ to the actuator so that the difference between (a) and (b) becomes zero. The feedback compensator FB is a feedback compensator for the pseudo posture angle of the vehicle body, and can be designed by PID control theory, phase lead / lag compensation, H∞ compensation, and the like. For example, the feedback compensator FB can be a PID compensator. By adopting the configuration shown in FIG. 10, the gain of the feedback compensator FB can be set larger than in the case where the rotation angle (θ−θ _n ) of the output shaft of the actuator is not used, and the disturbance τ _dis Can be accurately controlled. In the configuration shown in FIG. 10, the disturbance observer according to the embodiment or the control loop for feeding back the pseudo posture angular velocity shown in FIG. 9 may be incorporated inside the control loop for feeding back the pseudo posture angle.

  The technical elements described in this specification or the drawings exhibit technical usefulness alone or in various combinations, and are not limited to the combinations described in the claims at the time of filing. In addition, the technology illustrated in the present specification or the drawings achieves a plurality of objects at the same time, and has technical utility by achieving one of the objects.

DESCRIPTION OF SYMBOLS 10 Car body 12a, 12b Front wheel 16 Gyro sensor 18 Actuator 20 Control apparatus

Claims (4)

A wheel type moving body that drives a wheel and travels on a road surface,
The car body,
Means for detecting or estimating a posture angle around the motion axis of the vehicle body;
An actuator that can adjust the attitude angle around the motion axis of the vehicle body,
Means for detecting the rotation angle of the output shaft of the actuator;
Using the first control value obtained by processing the detected or estimated posture angle of the vehicle body with a low-pass filter and the second control value obtained by processing the detected rotation angle of the output shaft of the actuator with a high-pass filter, disturbance generated in the actuator is reduced. And a control device having a disturbance compensation unit for compensation .   The wheel according to claim 1, wherein the control device includes a feedback compensation unit that performs feedback compensation on the actuator based on the detected or estimated posture angle of the vehicle body so that the posture angle of the vehicle body becomes a target posture angle. Type moving body. Disturbance compensation module includes a first control value and second-order differential value of the sum of the second control value, the difference between the control input to the actuator, to compensate a disturbance occurs in the actuator, the wheel according to claim 1 or 2 Type moving body.   The wheel type moving body according to any one of claims 1 to 3, wherein a breakpoint frequency of the low-pass filter is lower than a natural vibration frequency of the vehicle body, and a breakpoint frequency of the high-pass filter is lower than a natural vibration frequency of the vehicle body.

Images