JP2014008886A - Vehicle stability control device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To further improve vehicle stability when a vehicle turns in a three-wheeled vehicle having a position of center of gravity deviated from a vehicle body center line.SOLUTION: Acceleration/deceleration of a vehicle is detected on the basis of a change in engine torque Eg during straight traveling, and a corrected steering angle Str2 corresponding to corrected steering amount is calculated on the basis of a steering angle corresponding to steering amount during acceleration. After a target yaw rate Yr corresponding to a reference turning state is estimated on the basis of the steering angle and the corrected steering angle Str2 during turning time or the like, engine control amount and brake control amount are set on the basis of an actual yaw rate Yr corresponding to an actual turning state and the target yaw rate Yr to thereby control acceleration/deceleration of the vehicle.

Description

  The present invention is applied to a three-wheeled vehicle in which the center of gravity deviates from the line passing through both wheels of a motorcycle that generates driving and braking force, such as a three-wheeled vehicle such as a motorcycle with a sidecar, and the vehicle using a vehicle brake control device or the like. The present invention relates to a vehicle stability control device that controls stability.

  In a three-wheeled vehicle such as a motorcycle with a sidecar, the position of the center of gravity is shifted from a line passing through both wheels of the motorcycle that generates driving and braking force (hereinafter referred to as a vehicle body centerline). That is, in a three-wheeled vehicle such as a motorcycle with a sidecar, the wheels on the motorcycle side are laterally deviated from the center of gravity. Further, the sidecar side wheel is merely a driven wheel. For this reason, the three-wheeled vehicle has a unique driving characteristic in which the turning characteristic is completely different between the turning to the motorcycle side and the turning to the side car side as the characteristic when turning the vehicle. Therefore, a high level of skill is required for driving such a three-wheeled vehicle, and a high level of skill is required to perform an appropriate operation during limit driving.

  In contrast, in Patent Document 1, conventionally, a sidecar brake is provided on a wheel on the sidecar side, and a braking force can be generated also on the wheel on the sidecar side based on a brake operation on the bike side. A brake device is disclosed that suppresses the occurrence of the above and compensates for stability during deceleration.

JP 2007-91087 A

  However, it is not possible to obtain a sufficient effect with respect to vehicle stability when the vehicle is turning only by providing a sidecar brake on the sidecar side wheel. Further, it is necessary to provide a special configuration called a sidecar brake on the sidecar side, which increases the complexity of the device and increases the cost associated with an increase in the number of parts.

  In this example, a motorcycle with a sidecar is given as an example of a three-wheeled vehicle in which the position of the center of gravity is shifted from the center line of the vehicle body. A similar problem occurs.

  In view of the above, the present invention provides a vehicle stability control device that can further improve vehicle stability during vehicle turning in a three-wheeled vehicle in which the position of the center of gravity is shifted from the vehicle body center line. And Further, in a three-wheeled vehicle in which the position of the center of gravity is shifted from the vehicle body center line, a vehicle stability control device that can further improve the vehicle stability when turning the vehicle without providing a brake mechanism on the side wheel side is provided. This is the second purpose.

  In order to achieve the above object, in the vehicle stability control device according to claim 1, the change in the drive torque and the steering amount when the change amount acquisition means (120) detects acceleration / deceleration of the vehicle during straight traveling. Change of the corrected steering amount with respect to the change of the drive torque is acquired based on the change of the vehicle torque, the reference turning state estimation means (125) estimates the reference turning state of the vehicle, and the vehicle stabilization control means (135, 145). ), The acceleration / deceleration of the vehicle is controlled by controlling the braking means (1) for generating a braking force on at least one of the driven wheel and the driving wheel based on the actual turning state and the reference turning state. In such a vehicle stability control device, the reference turning state estimating means (125) has a corrected steering amount estimating means (215) for estimating a corrected steering amount by the driver, which is acquired by the steering amount acquiring means (105). The reference turning state is estimated based on the steering amount and the corrected steering amount estimated by the corrected steering amount estimating means (215), and the vehicle stabilization control means (135, 145) is the actual turning state acquisition means (105). The acceleration / deceleration of the vehicle is controlled based on the acquired actual turning state, the reference turning state, and the change of the corrected steering amount with respect to the change of the driving torque acquired by the change amount means (120).

  Thus, when acceleration / deceleration of the vehicle is detected during straight traveling, a change in the corrected steering amount with respect to a change in drive torque is acquired. Then, the reference turning state is estimated based on the steering amount and the corrected steering amount, and the acceleration / deceleration of the vehicle is controlled based on the actual turning state, the reference turning state, and the change in the corrected steering amount with respect to the change in drive torque. Thereby, the vehicle stability at the time of vehicle turning in the three-wheeled vehicle in which the position of the center of gravity is shifted from the vehicle body center line can be further improved.

  For example, as described in claim 2, the vehicle stabilization control means (135, 145) is an oversteer state in which the vehicle is turning toward the side wheel and the turning degree in the actual turning state is greater than the reference turning state. , The acceleration / deceleration is controlled so that the acceleration of the vehicle decreases or the deceleration increases. As a result, the oversteer state can be suppressed, and the vehicle stability during vehicle turning can be further improved.

  Furthermore, as described in claim 3, the vehicle is provided with a lateral acceleration acquisition means for acquiring a lateral acceleration applied to the vehicle, and the vehicle stabilization control means (135, 145) When the lateral acceleration acquired by the acceleration acquisition means is equal to or greater than a threshold value that is a condition for lifting the side wheels, the acceleration / deceleration can be controlled so that the acceleration of the vehicle decreases or the deceleration increases.

  As described above, when the vehicle is turning to the side wheel side and the lateral acceleration acquired by the lateral acceleration acquisition means is equal to or higher than a threshold value that is a condition for lifting the side wheel, the acceleration / deceleration is controlled to control the side wheel. Can be prevented from lifting. Thereby, the vehicle stability at the time of vehicle turning can be improved more.

  Further, as described in claim 4, the vehicle stabilization control means (135, 145) is such that the vehicle is turning to the side opposite to the side wheels, and the turning degree in the actual turning state is smaller than the reference turning state. When the vehicle is in the understeer state, the acceleration / deceleration is controlled so that the acceleration of the vehicle decreases or the deceleration increases. Thereby, it becomes possible to suppress an understeer state, and the vehicle stability at the time of vehicle turning can be improved more.

  Further, as described in claim 5, the vehicle stabilization control means (135, 145) allows the vehicle deceleration to be within a predetermined limit deceleration while the vehicle is turning to the side opposite to the side wheels. It can also be controlled. As a result, it is possible to suppress the lifting of the rear wheels of the vehicle, and the vehicle stability when the vehicle turns can be further improved.

  In this case, as described in claim 6, the vehicle stabilization control means (135, 145) restricts as the lateral acceleration acquired by the lateral acceleration acquisition means increases while the vehicle is turning to the side wheel. It is preferable to set the deceleration small. As the lateral acceleration increases, the rear wheel lifts even with a smaller deceleration. Therefore, it is preferable to set the limited deceleration smaller as the lateral acceleration increases.

The invention according to claim 7 is characterized in that the side wheel is a driven wheel, and the side wheel is not provided with a brake mechanism. As described above, even when the side wheels are driven wheels and the brake mechanism is not provided, the vehicle can be stabilized by executing the control described in the above claims. Therefore, in a three-wheeled vehicle in which the position of the center of gravity deviates from the vehicle body center line, the vehicle stability during vehicle turning can be further improved without providing a brake mechanism on the side wheel side. The reference numeral indicates an example of the correspondence relationship with the specific means described in the embodiment described later.

It is the figure which showed the whole structure of the vehicle brake control apparatus 1 applied to the three-wheeled vehicle comprised with the motorcycle with a sidecar concerning 1st Embodiment of this invention. (A)-(c) is the figure which showed the moment which generate | occur | produces at the time of driving | running | working in a three-wheeled vehicle. It is the flowchart which showed the detail of the vehicle stability control process which brake ECU4 performs. It is the map which showed an example of the relationship between the correction value H and the control amount correction coefficient Hk. It is the flowchart which showed the detail of the target yaw rate Yr estimation process. It is the map which showed an example of the relationship between vehicle body speed V0 and running resistance. It is the map which showed an example of the relationship between correction torque Ts and correction steering angle Str2. (A) is a figure explaining the mechanism of the lift of the wheel at the time of left turn, (b) is a figure explaining the mechanism of the lift of the wheel at the time of right turn. It is the model figure which showed the mode at the time of load movement of a three-wheeled vehicle.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings. In the following embodiments, parts that are the same or equivalent to each other will be described with the same reference numerals.

(First embodiment)
A first embodiment of the present invention will be described. In the present embodiment, a vehicle brake control device applied to a three-wheeled vehicle configured by a motorcycle with a side car provided with side wheels will be described as an example.

  FIG. 1 shows an overall configuration of a vehicle brake control device 1 applied to a three-wheeled vehicle constituted by a motorcycle with a sidecar. This vehicle brake control device 1 applies a brake to the front wheel FW as a driven wheel and the rear wheel RW as a driving wheel, and controls the acceleration / deceleration of the vehicle by generating a braking force in the stability control. It is also used to improve the stability of the vehicle. Specifically, the vehicle brake control device 1 has a configuration including two piping systems, a system that generates a braking force for the front wheels FW and a system that generates a braking force for the rear wheels RW. .

  As shown in FIG. 1, the vehicle brake control device 1 includes a brake lever 11 positioned on the right side of the steering wheel and a brake pedal 12 positioned in front of the right footrest. The brake lever 11 and the brake pedal 12 correspond to brake operation members for generating a braking force for the front wheel FW and the rear wheel RW, respectively, and are operated independently by the driver. The brake lever 11 and the brake pedal 12 are connected to a brake circuit having first and second piping systems 14 and 15 via a front wheel master cylinder (hereinafter referred to as M / C) 13a and a rear wheel M / C 13b. Has been.

  The brake lever 11 is connected to a first piping system 14 that generates a braking force for the front wheel FW via the front wheel M / C 13a and the like. The brake pedal 12 is connected to a second piping system 15 that generates a braking force for the rear wheel RW via the rear wheel M / C 13b and the like. When the brake lever 11 or the brake pedal 12 is operated, the front wheel FW side is passed through the first and second piping systems 14 and 15 based on the brake fluid pressure generated in the front wheel M / C 13a and the rear wheel M / C 13b. W / C pressure is generated with respect to the front wheel cylinder (hereinafter referred to as W / C) 16 and the rear wheel W / C 17 on the rear wheel RW side, thereby generating a braking force.

  A brake fluid pressure control actuator 2 is provided between the front wheel M / C 13a and the rear wheel M / C 13b and the front wheel W / C 16 and the rear wheel W / C 17. The first piping system 14 controls the W / C pressure applied to the front wheels W / C 16 by controlling various control valves and the like provided in the brake fluid pressure control actuator 2, and the second piping system 15 controls the rear wheels. The W / C pressure applied to the W / C 17 is controlled.

  Hereinafter, although the detailed structure of the 1st, 2nd piping system 14 and 15 is demonstrated, since the 1st piping system 14 and the 2nd piping system 15 are the substantially the same structures, here about the 1st piping system 14 The second piping system 15 will be described with reference to the first piping system 14.

  The first piping system 14 is provided with a pipeline A serving as a main pipeline connecting the front wheel M / C 13a and the front wheel W / C16. Through this pipe A, the M / C pressure is transmitted to generate the W / C pressure at the front wheel W / C 16.

  Further, the pipe line A is provided with a first differential pressure control valve 18 that can be controlled between a communication state and a differential pressure state. The first differential pressure control valve 18 is in a communicating state in the normal brake state, and is in a differential pressure state when a current is passed through the solenoid. The differential pressure formed by the first differential pressure control valve 18 changes according to the current value of the current flowing through the solenoid, and the larger the current value, the larger the differential pressure amount. When the first differential pressure control valve 18 is in the differential pressure state, the flow of the brake fluid is regulated so that the W / C pressure is higher than the M / C pressure by the amount of the differential pressure.

  The pipe A has a first pressure increase control valve 19 that controls the increase of the brake fluid pressure to the front wheel W / C 16 on the downstream side of the first differential pressure control valve 18 on the front wheel W / C 16 side. Is provided. The 1st pressure increase control valve 19 is comprised by the solenoid valve as a 2 position valve which can control a communication and interruption | blocking state. When the first pressure increase control valve 19 is controlled to be in a communication state, the M / C pressure or the brake fluid pressure generated by the discharge of brake fluid from the pump 22 described later is applied to the front wheels W / C16.

  During normal braking by the driver operating the brake lever 11, the first differential pressure control valve 18 and the first pressure increase control valve 19 are constantly controlled. For this reason, the M / C pressure generated in the front wheel M / C 13a is directly transmitted as the W / C pressure of the front wheel W / C16. The first differential pressure control valve 18 and the first pressure increase control valve 19 are respectively provided with safety valves 18a and 19a in parallel.

  A pipe B as a pressure reducing pipe is connected between the first pressure increase control valve 19 and the front wheel W / C 16 in the pipe A, and a pressure regulating reservoir 20 is provided for the pipe B. Further, a first pressure reduction control valve 21 made of an electromagnetic valve is disposed in the pipeline B as a two-position valve that can control the communication / blocking state. The first pressure reducing control valve 21 is always cut off during normal braking.

  Further, a conduit C serving as a reflux conduit is disposed so as to connect the pressure regulating reservoir 20 and the first differential pressure control valve 18 and the first pressure increase control valve 19 in the conduit A. The pipe C is provided with a self-priming pump 22 driven by the motor 3 so as to suck and discharge brake fluid from the pressure regulating reservoir 20 toward the M / C 13 side or the front wheel W / C 16 side. .

  And the pipe line D used as an auxiliary pipe line is provided so that the pressure regulation reservoir 20 and the front wheel M / C13a may be connected. Brake fluid is sucked from the front wheel M / C 13a by the pump 22 through this pipe D and discharged to the pipe A through the pipe C, so that brake fluid is supplied to the front wheel W / C 16 side in anti-skid control or the like. In addition, the W / C pressure of the front wheel W / C 16 can be increased.

  The pressure regulating reservoir 20 is connected to the pipeline D to receive the brake fluid from the front wheel M / C 13a side, and the reservoir fluid 20a is connected to the pipeline B and the pipeline C and discharged from the front wheel W / C16. A reservoir hole 20b for receiving and supplying brake fluid to the suction side of the pump 22 is provided and communicates with the reservoir chamber 20c. A valve body 20d made of a ball valve is disposed inside the reservoir hole 20a. The valve body 20d is attached to and detached from the valve seat 20e to control the communication interruption between the pipe D and the reservoir chamber 20c, and the distance between the valve seat 20e and the reservoir chamber 20c is adjusted. The pressure difference between the internal pressure and the M / C pressure is regulated. Below the valve body 20d, a rod 20f having a predetermined stroke for moving the valve body 20d up and down is provided separately from the valve body 20d. Also, in the reservoir chamber 20c, there are a piston 20g interlocking with the rod 20f, and a spring 20h that generates a force for pressing the piston 20g toward the valve body 20d to push out the brake fluid in the reservoir chamber 20c. Is provided.

  The pressure regulating reservoir 20 configured as described above is configured such that when a predetermined amount of brake fluid is stored, the valve body 20d is seated on the valve seat 20e and the brake fluid does not flow into the pressure regulating reservoir 20. . Therefore, more brake fluid than the suction capacity of the pump 22 does not flow into the reservoir chamber 20c, and no high pressure is applied to the suction side of the pump 22.

  On the other hand, as described above, the second piping system 15 has substantially the same configuration as the first piping system 14. That is, the first differential pressure control valve 18 and the safety valve 18a correspond to the second differential pressure control valve 23 and the safety valve 23a. The first pressure increase control valve 19 and the safety valve 19a correspond to the second pressure increase control valve 24 and the safety valve 24a, respectively, and the first pressure reduction control valve 21 corresponds to the second pressure reduction control valve 26. The pressure regulating reservoir 20 and the respective components 20a to 20h correspond to the pressure regulating reservoir 25 and the respective components 25a to 25h. The pump 22 corresponds to the pump 27. Further, the pipeline A, the pipeline B, the pipeline C, and the pipeline D correspond to the pipeline E, the pipeline F, the pipeline G, and the pipeline H, respectively. The brake fluid pressure control actuator 2 is configured as described above. The various control valves 18, 19, 21, 23, 24, 26 and the motors 3 for driving the pumps 22, 27 provided in the brake fluid pressure control actuator 2 are electronic control devices for brake control ( (Hereinafter referred to as brake ECU) 4.

  The brake ECU 4 corresponds to the vehicle stability control device of the present invention that controls the control system of the vehicle brake control device 1, and is constituted by a known microcomputer including a CPU, a ROM, a RAM, an I / O, and the like. Then, processing such as various calculations is executed in accordance with a program stored in the ROM or the like. For example, the brake ECU 4 receives detection signals from the wheel speed sensor 5, the yaw rate sensor 6, and the steering angle sensor 7 provided on the front wheel FW and the rear wheel RW, and obtains various physical quantities.

  Specifically, the brake ECU 4 performs vehicle stability control in addition to anti-skid control and performs various calculations for executing these controls. Therefore, the brake ECU 4 calculates, for example, the wheel speed of each wheel FW and RW, the vehicle body speed (estimated vehicle speed), the slip ratio of each wheel, etc. based on the detection signal of the wheel speed sensor 5. Further, the brake ECU 4 obtains the yaw rate Yr, the steering angle sensor value Str, and the like based on the detection signals of the yaw rate sensor 6 and the steering angle sensor 7. In the case of the present embodiment, for the yaw rate Yr, a positive value means a counterclockwise moment with respect to the vehicle rotation center, a value generated when turning left, etc., and a negative value means a clockwise moment with respect to the vehicle rotation center. It is assumed that the value is generated when turning right. Further, it is assumed that the steering angle sensor value Str means a positive value when the steering angle is turned in the left turn direction and a negative value is turned in the right turn direction.

  The brake ECU 4 communicates information with the engine ECU 8 that performs engine control, and acquires information on the engine torque (drive torque) from the engine ECU 8. Further, the brake ECU 4 transmits information related to the engine control amount to the engine ECU 8. The engine ECU 8 adjusts the engine torque based on the information on the engine control amount so as to achieve vehicle stability.

  Then, the brake ECU 4 performs various controls such as anti-skid control and vehicle stability control based on various calculation results, determines a control target wheel when executing various controls, and determines a brake control amount for the control target wheel. And obtain the engine control amount. Based on the result, the brake ECU 4 controls the supply current to the various control valves 18, 19, 21, 23, 24, 26 and the motor 3, and transmits information on the engine control amount to the engine ECU 8. With the configuration as described above, the vehicle brake control device 1 according to the present embodiment is configured.

  In the vehicle brake control device 1 configured as described above, for example, during normal braking in which anti-skid control or the like is not performed, the brake ECU 4 sends various control valves 18, 19, 21, 23, 24, 26 and the motor 3. No current is supplied. For this reason, the W / C pressure corresponding to the operation amount of the brake lever 11 and the brake pedal 12 is generated in each of the W / Cs 16 and 17. As a result, a braking force corresponding to the brake lever 11 and the brake pedal 12 is generated on the front wheel FW and the rear wheel RW.

  Further, during anti-skid control, current is supplied from the brake ECU 4 to the various control valves 18, 19, 21, 23, 24, 26 and the motor 3 as necessary. As a result, the pipelines A and E and the pressure regulating reservoirs 20 and 25 are brought into communication with each other through the pipelines B and F, the W / C pressure generated in each of the W / Cs 16 and 17 is reduced, and the wheel slip is reduced. It becomes possible to avoid wheel lock by being suppressed.

  During vehicle stability control, a W / C pressure corresponding to the brake control amount is generated to generate a braking force on the wheel to be controlled, or information on the engine control amount is transmitted to the engine ECU 8. . Since this vehicle stability control is a characteristic feature of the present invention, the vehicle stability control will be described in detail below.

  Vehicle stability control in this embodiment is vehicle control that improves turning performance for a three-wheeled vehicle in which the center of gravity deviates from the vehicle body center line passing through both wheels FW and RW of the motorcycle, such as a three-wheeled vehicle such as a motorcycle with a sidecar. It shows that. First, moments generated in a three-wheeled vehicle like this embodiment will be described.

  FIGS. 2A to 2C are views showing moments generated during traveling in the three-wheeled vehicle of the present embodiment. FIG. 2A is a straight traveling, FIG. 2B is a left turn, and FIG. It shows the moment generated when turning right.

  As shown in FIGS. 2A to 2C, when the three-wheeled vehicle of the present embodiment is provided with a side car 31 on the left side of the motorcycle 30, the vehicle body that passes through both wheels FW and RW of the motorcycle 30. The center of gravity shifts to the left with respect to the center line.

  Therefore, during straight traveling, when a driving force is generated at the rear wheel RW as shown in FIG. 2A, a moment around the center of gravity, that is, a counterclockwise moment is generated based on the driving force. In order to suppress such a moment, it is necessary to generate a clockwise moment. Usually, the driver travels straight by performing correction steering with respect to the moment.

  Thus, a counterclockwise moment is originally generated with respect to the center of gravity. For this reason, when performing a left turn, which is a turn on the side car 31 side, the turning degree is larger than the ideal turning state corresponding to the steering operation of the driver, and the vehicle is in an oversteer state where the vehicle cuts inwardly. Can become unstable. Also, when making a right turn, which is a turn on the side of the motorcycle 30 opposite to the side car 31 side, the turning degree is smaller than the ideal turning state corresponding to the driver's steering operation, and the vehicle protrudes outward. Understeer, and the vehicle can become unstable.

  Therefore, in the case of a left turn, as shown in FIG. 2 (b), the front and rear wheels FW, RW or either one of the front and rear wheels FW, RW are controlled so as to generate a deceleration to generate a clockwise moment. Generate power or reduce engine torque. Thereby, it becomes possible to suppress an oversteer state. Also in the case of a right turn, as shown in FIG. 2 (c), in order to generate a deceleration to the wheels in order to generate a clockwise moment, the front and rear wheels FW, RW or either one of them Generate braking force or reduce engine torque. Thereby, it becomes possible to suppress an understeer state. In this way, it is possible to suppress an oversteer state in a left turn and an understeer state in a right turn, and to realize vehicle stability that improves turning performance.

  Next, details of wheel stability control executed based on the above mechanism will be described. FIG. 3 is a flowchart showing details of the vehicle stability control process executed by the brake ECU 4. The processing shown in this figure is executed at predetermined control cycles when, for example, an ignition switch (not shown) is on.

  First, in step 100, as an initial setting, the engine control amount is set to 0, the front brake control amount is set to 0, and the rear brake control amount is set to 0. The engine control amount is a control amount corresponding to the adjustment amount of the engine torque, and is transmitted to the engine ECU 8. The front brake control amount and the rear brake control amount mean control amounts of braking force generated on the front wheels FW and the rear wheels RW.

  Here, with respect to the engine control amount, engine control amount = 0 means that nothing is adjusted, and when the engine control amount is generated, it means that the engine torque is decreased by that amount. For the front brake control amount and the rear brake control amount, brake control amount = 0 means that nothing is adjusted for the brake. When the brake control amount is generated, a predetermined differential pressure (for example, 10 MPa) is generated by the first and second differential pressure control valves 18 and 23 and the motor 3 is operated to drive the pumps 22 and 27. This means that the brake is applied by increasing the W / C pressure of the front wheel W / C16 and the rear wheel W / C17. Therefore, by setting each control amount to 0 as an initial setting in step 100, no engine torque or brake is adjusted.

  Subsequently, the routine proceeds to step 105, where the yaw rate Yr and the steering angle sensor value Str are detected based on the detection signals of the yaw rate sensor 6 and the steering angle sensor 7, and the engine torque Eg is acquired from the engine ECU.

  Then, the process proceeds to step 110 where the yaw rates Y (n−1) and Y (n) obtained in the previous and current control cycles are both 0, and the engine torque Eg (n obtained in the current and previous control cycles is obtained. ) And Eg (n−1) are not equal (Eg (n) ≠ Eg (n−1)). A situation in which both the yaw rates Y (n−1) and Y (n) obtained in the previous and current control cycles are 0 means that the vehicle is traveling straight ahead. Also, the situation where the engine torques Eg (n) and Eg (n-1) obtained in the current and previous control cycles are not equal is that the engine torque Eg has fluctuated and the accelerator has been opened and closed, that is, applied to the vehicle. It means that there was a slowdown. In such a situation, it is considered that the driver maintains the straight traveling by performing the correction steering for the moment.

Therefore, if an affirmative determination is made in this step, the process proceeds to steps 115 and 120 to perform a calculation process of the control amount correction coefficient Hk. If a negative determination is made, the process proceeds to step 125 without proceeding to steps 115 and 120. In step 115, first, a correction value H is calculated. The correction value H is a correction value resulting from the movement of the center of gravity due to loading on the sidecar 31 or the like. In this embodiment, the correction value H is obtained by calculating the yaw rate change due to the change in the center of gravity position on the desk. Specifically, as shown in step 115, the correction value H is calculated by the following equation. In the following equation, S2 is a coefficient determined for each vehicle based on vehicle characteristics, and Str (n) and Str (n-1) are the steering angle sensor values Str in the current and previous control cycles.

(Equation 1)
H = S2 * {Str (n) -Str (n-1)} / {Eg (n) -Eg (n-1)}
In this equation, the difference in the steering angle St (Str (n) −Str (n−1)) between the current control cycle and the previous control cycle corresponds to the driver's corrected steering amount performed during straight traveling. In addition, the difference (Eg (n) −Eg (n−1)) between the engine torque Eg between the current control cycle and the previous control cycle corresponds to the fluctuation amount of the engine torque Eg during straight traveling. Then, the value obtained by dividing the corrected steering amount by the fluctuation amount of the engine torque Eg is the corrected steering amount corresponding to the center of gravity position. Therefore, by multiplying this value by the coefficient S2, the correction value H corresponding to the change in the center of gravity position is obtained. Can be calculated.

  Subsequently, in step 120, the control amount correction coefficient Hk is calculated based on the correction value H calculated in step 115. The control amount correction coefficient Hk is a coefficient used for correcting the engine control amount and the brake control amount corresponding to the correction value H.

  The correction value H is a large value in a situation where the corrected steering amount increases when a slight engine torque change occurs. In this situation, the yaw rate changes greatly even with a slight change in engine torque. In such a case, when the yaw rate is to be generated so as to suppress the moment, a large yaw rate can be generated by a change in the small engine control amount or the brake control amount. A large yaw rate can be generated even if changes in the engine control amount and the brake control amount are small. Therefore, the control amount correction coefficient Hk corresponding to the correction value H can be calculated based on such a relationship that the control amount correction coefficient Hk becomes smaller as the correction value H becomes larger. FIG. 4 is a map showing an example of the relationship between the correction value H and the control amount correction coefficient Hk. As shown in this figure, the control amount correction coefficient Hk corresponding to the correction value H can be calculated based on such a relationship that the control amount correction coefficient Hk gradually decreases as the correction value H increases.

  After calculating the control amount correction coefficient Hk in this way, the routine proceeds to step 125 where target yaw rate Yr estimation processing is performed. The target yaw rate Yr is a yaw rate Yr assumed in an ideal turning state, that is, a reference turning state. Details of this processing will be described with reference to a flowchart showing details of target yaw rate Yr estimation processing in FIG.

  During steady (constant speed) driving, the steering angle is almost in the position corresponding to the course, but during acceleration / deceleration, the magnitude of the acceleration / deceleration is required to correct the course against the yaw moment generated according to the acceleration / deceleration. Corrective steering according to the need is required. In order to obtain the target yaw rate Yr, it is necessary to eliminate the corrected steering angle Str2 and obtain a reference steering angle that is a steering angle corresponding to the course.

  Basically, since “torque / vehicle weight” corresponds to acceleration, acceleration is obtained from engine torque Eg during acceleration. However, the engine torque Eg is used not only for acceleration but also for overcoming the running resistance and maintaining the vehicle speed. Therefore, by subtracting the torque corresponding to the running resistance from the engine torque Eg, a correction torque Ts that is an amount contributing to the acceleration of the engine torque Eg is calculated, and the corrected steering angle Str2 is calculated from the correction torque Ts. The target yaw rate Yr must be estimated by calculating and further obtaining the reference steering angle. For this reason, in steps 200 to 220, the target yaw rate Yr is estimated in the following order.

  First, at step 200, the vehicle body speed V0 used for running resistance calculation is detected. The vehicle body speed V0 is calculated by a known method using the calculation result after calculating the wheel speed of each wheel FW and RW based on the detection signal of the wheel speed sensor 5.

  Subsequently, the routine proceeds to step 205, where the running resistance torque Td is calculated based on the vehicle body speed V0 detected at step 200. The running resistance torque Td is a value obtained by converting a running resistance applied to the vehicle during running, and the vehicle body speed V0 and the running resistance have a certain relationship. For this reason, for example, based on design values and experimental value settings, the running resistance torque Td corresponding to the running resistance for each vehicle speed V0 is mapped in advance, and the running resistance torque Td is calculated from the detected vehicle speed V0 and the map. can do.

  FIG. 6 is a map showing an example of the relationship between the vehicle speed V0 and the running resistance. As shown in this figure, as a map in which the traveling resistance torque Td increases as the vehicle body speed V0 increases and the degree of increase in the traveling resistance torque Td increases as the vehicle body speed V0 increases, The relationship with running resistance can be expressed. Note that the travel resistance is not actually determined only by the vehicle body speed V0, but is also affected by the inclination of the travel road surface. Therefore, the travel resistance torque Td may be corrected based on the slope of the travel road surface as necessary. .

  Then, the process proceeds to step 210, and the correction torque Ts is calculated. The correction torque Ts is calculated by subtracting the running resistance torque Td from the engine torque Eg (n) in the current control cycle. Thus, by subtracting the running resistance equivalent torque Td from the engine torque Eg (n), the correction torque Ts corresponding to the torque component contributing to the acceleration in the engine torque Eg (n) can be extracted.

  Thereafter, the process proceeds to step 215, and a corrected steering angle Str2 is calculated based on the correction torque Ts. Since the correction torque Ts, that is, the torque component of the engine torque Eg that contributes to acceleration is a value corresponding to the correction steering, the correction steering Str2 can be calculated based on the correction torque Ts. For example, the correction rudder angle amount corresponding to the correction torque Ts can be mapped in advance based on the design value or the experimental value, and the correction rudder angle Str2 can be calculated from the calculated correction torque Ts and this map.

  FIG. 7 is a map showing an example of the relationship between the correction torque Ts and the corrected steering angle Str2. As shown in this figure, the corrected steering angle Str2 has a value corresponding to the magnitude of the correction torque Ts, and the corrected steering angle Str2 increases as the correction torque Ts increases. Specifically, with respect to the correction torque Ts, the sign is reversed depending on whether it is on the acceleration side or the deceleration side. Therefore, the absolute value of the correction steering angle Str2 increases as the absolute value of the correction torque Ts increases. Note that even when the correction torque Ts is the same value with a change in the vehicle weight, that is, a change in the position of the center of gravity, the moment for changing the vehicle changes, and the corrected steering angle Str2 also changes. For this reason, the relationship between the correction torque Ts and the necessary corrected steering angle Str2 may be corrected based on the control amount correction coefficient Hk that represents the yaw rate generation tendency obtained in advance. For example, as shown by a solid line and a broken line in FIG. 7, when the control amount correction coefficient Hk is small, the inclination of the change in the modified steering angle Str2 with respect to the correction torque Ts is large, and when the control amount correction coefficient Hk is large. You may make it the inclination become small.

  Finally, the routine proceeds to step 220, where the steering angle sensor value Str obtained from the detection signal of the steering angle sensor 7, that is, the correction steering angle Str2 is subtracted from the steering angle that is actually generated, so that the steering angle corresponding to the course is obtained. Calculate a certain reference rudder angle. Then, the target yaw rate Yr is estimated based on the reference steering angle and the vehicle body speed V0. This estimation method is the same as that for a four-wheeled vehicle, and the target yaw rate Yr to be generated according to the course can be estimated from the turning radius obtained from the reference rudder angle and the vehicle body speed V0. In this way, the target yaw rate Yr estimation process is completed.

  Then, after estimating the target yaw rate Yr, the process proceeds to step 130 in FIG. 3, and the actual yaw rate Yr actually generated currently obtained from the current steering angle sensor value Str and the detection signal of the yaw rate sensor 6 is positive, In addition, it is determined whether or not the actual yaw rate Yr exceeds the target yaw rate Yr. Here, if the rudder angle sensor value Str and the actual yaw rate Yr are positive, it indicates a state of turning left, that is, turning to the side car 31 side, and the actual yaw rate Yr exceeds the target yaw rate Yr. This means that the vehicle is in an oversteer state.

  Therefore, if an affirmative determination is made in this step, the routine proceeds to step 135, where an engine control amount, a front brake control amount, or a rear brake control amount for suppressing the oversteer state is calculated. Specifically, the absolute value of the value obtained by subtracting the target yaw rate Yr from the actual yaw rate Yr corresponds to the yaw rate error. Therefore, the engine control amount can be calculated by multiplying the absolute value of the difference between the actual yaw rate Yr and the target yaw rate Yr by the constant E1 and the control amount correction coefficient Hk. Further, the front brake control amount or the rear brake control amount can be calculated by multiplying the absolute value of the difference between the actual yaw rate Yr and the target yaw rate Yr by the constant B1 and the control amount correction coefficient Hk. In this way, the engine control amount, front brake control amount, or rear brake control amount for suppressing the oversteer state can be calculated.

  If a negative determination is made in step 130, the process proceeds to step 140, where the actual yaw rate Yr actually generated currently obtained from the current steering angle sensor value Str or the detection signal of the yaw rate sensor 6 is negative, and Then, it is determined whether the actual yaw rate Yr is less than the target yaw rate Yr. Here, if the steering angle sensor value Str and the actual yaw rate Yr are negative, it indicates a state of turning right, that is, turning to the motorcycle 30 side, and the actual yaw rate Yr is less than the target yaw rate Yr. If it exists, it means understeer.

  Therefore, if an affirmative determination is made in this step, the routine proceeds to step 145, where an engine control amount, a front brake control amount, or a rear brake control amount for suppressing the understeer state is calculated. Specifically, the engine control amount, the front brake control amount, or the rear brake control amount for suppressing the understeer state can be calculated by the same method as in step 135 described above.

  After calculating the engine control amount, the front brake control amount, or the rear brake control amount in this way, various processes are performed based on the results. That is, regarding the engine control amount, information indicating the engine control amount is transmitted to the engine ECU 8. As a result, the engine ECU 8 reduces the driving force by subtracting the engine control amount from the engine torque Eg. For each brake control amount, various components of the brake fluid pressure control actuator 2 are driven so that a corresponding braking force is generated. Through these processes, acceleration is reduced or deceleration is increased, and the acceleration / deceleration of the vehicle can be adjusted to generate a clockwise moment, suppressing oversteer and understeer conditions. It becomes possible.

  As described above, in the present embodiment, the vehicle acceleration / deceleration is detected based on the change in the engine torque Eg during straight traveling, and the correction corresponding to the correction steering amount is performed based on the steering angle corresponding to the steering amount during acceleration. Calculate the rudder angle. Then, at the time of turning, the target yaw rate Yr corresponding to the reference turning state is estimated based on the steering angle and the corrected steering angle Str2, and then based on the actual yaw rate Yr corresponding to the actual turning state and the target yaw rate Yr. The vehicle acceleration / deceleration is controlled by setting the engine control amount and the brake control amount.

  Thereby, in the three-wheeled vehicle in which the position of the center of gravity is shifted from the vehicle body center line, the turning performance can be improved, and the vehicle stability at the time of turning the vehicle can be further improved. Specifically, it is possible to suppress an oversteer state or an understeer state, and the vehicle stability during vehicle turning can be further improved. Such control can be performed without the brake mechanism provided on the sidecar 31, and therefore, in a three-wheeled vehicle in which the position of the center of gravity is deviated from the center line of the vehicle body, the brake mechanism is not provided on the sidecar 31 side. This makes it possible to further improve the vehicle stability when the vehicle turns.

(Second Embodiment)
A second embodiment of the present invention will be described. In this embodiment, the lifting of the wheel can be suppressed as compared with the first embodiment, and the other parts are the same as those in the first embodiment. Therefore, only the parts different from the first embodiment will be described.

  In the present embodiment, as the vehicle stability control, in addition to improving the turning performance of the three-wheeled vehicle whose center of gravity is shifted from the line passing through both wheels of the motorcycle, control for suppressing the lifting of the wheel is also performed.

  As described in the first embodiment, the vehicle becomes unstable even in the oversteer state or the understeer state, but the vehicle may become unstable due to the lifting of the wheels during turning. In the present embodiment, in addition to suppressing the oversteer state and the understeer state, the vehicle is also prevented from becoming unstable due to the lifting of the wheels.

  First, the lifting of the wheel will be described with reference to FIG. FIG. 8A is a diagram for explaining a mechanism for lifting a wheel when turning left, and FIG. 8B is a diagram for explaining a mechanism for lifting a wheel when turning right.

  As shown in FIG. 8A, when performing a left turn on the side car 31 side, lateral acceleration (hereinafter referred to as acceleration G) is generated in the right direction with respect to the vehicle body. Due to the lateral G, the vehicle rolls outward and the side wheels SW provided in the side car 31 are lifted. A three-wheeled vehicle such as the motorcycle 30 with the sidecar 31 has a high center of gravity position, and since the position of the centerline of the vehicle body connecting the front and rear wheels FW and RW of the motorcycle 30 and the center of gravity is close, the side wheel SW of the sidecar 31 is Lifts easily. On the other hand, as shown in FIG. 8B, when performing a right turn on the side of the motorcycle 30, a lateral G is generated in the left direction with respect to the vehicle body, and this and a deceleration generated by the right turn braking. G rolls the vehicle diagonally forward on the outside of the turn. For this reason, if the rear wheel RW is lifted and the lift is large, the lower surface of the side car 31 may be grounded, and the vehicle behavior may change suddenly. A three-wheeled vehicle such as the motorcycle 30 with the sidecar 31 has a high center of gravity, and the rear wheel RW is easily lifted because the center line of the vehicle body connecting the front and rear wheels FW and RW of the motorcycle 30 is close to the center of gravity. . In these cases, the stability of the vehicle cannot be maintained.

  Therefore, when the side wheel SW of the sidecar 31 is lifted in the left turn, the braking force is generated by generating a braking force for either or both of the front and rear wheels FW and RW or by reducing the engine torque. Is generated. Thereby, it becomes possible to suppress the lifting of the side wheel SW. Even when the rear wheel RW is lifted in a right turn, the braking force is generated by generating a braking force for both or either of the front and rear wheels FW, RW or by reducing the engine torque. Is generated. Thereby, it becomes possible to suppress the lifting of the rear wheel RW. In this way, it is possible to suppress the lifting of the wheel when turning left or turning right, and vehicle stabilization can be realized.

  Such a phenomenon does not always occur at the same time as the oversteer state or the understeer state, but also occurs separately. For this reason, what is necessary is just to generate | occur | produce the control amount for suppressing the lifting of a wheel separately from the control amount for suppressing an oversteer state and an understeer state. Therefore, if it occurs at the same time as the oversteer state or the understeer state, a control amount for suppressing the lifting of the wheel may be generated in addition to the control amount for suppressing the oversteer state or the understeer state.

  Conditions for lifting the side wheel SW of the side car 31 when turning left and conditions for raising the rear wheel RW of the bike 30 when turning right are obtained as follows. Each of these conditions will be described with reference to a diagram showing the state of the three-wheeled vehicle at the time of load movement in FIG.

  As shown in FIG. 9A, when the vehicle weight is W, the height to the center of gravity is h, and the lateral G generated in the three-wheeled vehicle is αy, the moment amount A applied to the rotation center of the three-wheeled vehicle is It is expressed as W × αy × h. When the tread (the arm length from the line connecting the wheels of the motorcycle 30 to the side wheel SW of the sidecar 31) is T, if the moment amount A is divided by the tread T, the load reduction amount acting at the side wheel position ΔW. That is, (W × αy × h) / T is the load reduction amount ΔW at the side wheel position. When the condition that the load reduction amount ΔW exceeds the load applied to the side wheel SW (side wheel load) is satisfied, the side wheel SW is lifted.

  The load moment B of the side wheel SW applied to the center of rotation of the three-wheeled vehicle (the longitudinal axis of the motorcycle 30) is the portion of the vehicle weight related to the side wheel SW × gravity × tread, that is, W × β × g (= 1) × T. β is a load coefficient of the side wheel SW and takes a value in a range of (0 <β <1). The load coefficient β is a distance TS between the center of gravity position CP on the tread and the side wheel SW obtained based on the vehicle specifications, and a distance TB between the center of gravity position CP on the tread and the line connecting the wheels of the motorcycle 30. It can be obtained from the ratio. Or you may derive from the result of having actually measured the load of each wheel using vehicles beforehand. Therefore, αy ≧ T × β / h can be derived from the following equation, and when the lateral G (= αy) exceeds a threshold value (= T × β / h) that is a condition for lifting the side wheel SW, it is over As in the steering state, by increasing the acceleration or increasing the deceleration, it is possible to suppress the lifting of the side wheels SW, and it is possible to further improve the stability of the vehicle.

(Equation 2)
(W × αy × h) ≧ (W × β × T)
(W × αy × h) ≧ W × (T × β)
αy × h ≧ T × β
αy ≧ T × β / h
On the other hand, when turning right, the load moment By of the rear wheel RW by the lateral G can be calculated when the model of FIG. At this time, the load (W × (1−β) × g × T) on the motorcycle 30 side is supported by the front and rear wheels FW and RW of the motorcycle 30. The rear wheel load is a value obtained by sharing the motorcycle side load with the front wheel FW. The rear wheel load can be calculated from the position of the center of gravity on the wheel base and the lengths of the front and rear wheels FW and RW, similarly to the load of the side wheel SW. On the other hand, since the moment due to the lateral G is also applied to the front and rear wheels FW and RW, the load moment of the rear wheel RW becomes By / 2 because the load moment By applied to the bike 30 side is shared by the front and rear wheels FW and RW. The load moment By is expressed by W × αy × h as in the case of turning to the side wheel SW side. The load reduction amount ΔWy acting on the rear wheel RW due to the lateral G is ½ of (W × αy × h) / T.

  Further, the lifting of the rear wheel RW of the motorcycle 30 due to the deceleration G corresponds to a case where the relationship between the wheels of the motorcycle 30 and the side wheels SW of the sidecar 31 is replaced with the front and rear wheels FW and RW of the motorcycle 30. For this reason, as shown in FIG. 9B, the tread T is replaced with the wheel base l and the lateral G ( = Αy) may be replaced with the deceleration G. That is, the moment amount Bx of the rear wheel RW due to the deceleration G is W × αx × h. On the other hand, the moment Bx due to the deceleration G is also Bx / 2 because the rear wheel RW and the side wheel SW are shared. The load reduction amount ΔWx acting at the rear wheel position by the deceleration G is ½ of (W × αx × h) / l.

  Therefore, for the load reduction amount acting on the rear wheel RW due to the lateral G and the deceleration G, the load reduction amount ΔWy acting on the rear wheel RW due to the lateral G and the load reduction amount ΔWx acting on the rear wheel RW due to the deceleration G are added. Value. Therefore, as shown in the following equation, when the condition that the load reduction amount acting on the rear wheel RW by the lateral G and the deceleration G becomes equal to or greater than the rear wheel load, the rear wheel RW is lifted.

(Equation 3)
((W × αx × h) / l) / 2 + ((W × αy × h) / T) / 2 ≧ rear wheel load Therefore, by adjusting the front brake control amount, the lateral G ( = Gy) and deceleration G (Gx) are suppressed so that the condition of Equation 3 is not satisfied. For example, a control limit amount map of αx and αy is prepared, and the control limit amount map is followed, or according to the control limit amount determined by simply multiplying αx and αy by a constant, the control limit amount is less than or equal to Set the front brake control amount to limit the deceleration so that it is less than or equal to the limited deceleration. Thereby, it is possible to suppress the speed of the vehicle while suppressing the lifting of the rear wheel RW, and it is possible to further improve the stability of the vehicle.

  When the moment acts on the two wheels, the moment acting on each wheel is halved. However, in the dynamic operation, there is a suspension operation or the like, and the moment does not necessarily act equally on the two wheels. . Therefore, ½ in the longitudinal and lateral moment calculation formulas may be corrected as necessary from the experimental results.

  In addition, since the rear wheel RW is more likely to be lifted as the lateral G is larger, it is preferable to set the control limit amount so that the smaller the lateral G is, the smaller the limit deceleration is. As a result, the rear wheel RW can be prevented from lifting regardless of the size of the lateral G.

(Other embodiments)
In the embodiment described above, the motorcycle 30 with the sidecar 31 is described as an example of the three-wheeled vehicle whose center of gravity is shifted from the vehicle body center line passing through both wheels FW and RW of the motorcycle 30 that generates the driving and braking force. It is not necessary to have a configuration provided with a motorcycle, and it may simply be a motorcycle provided with a side wheel SW. In the above embodiment, the motorcycle with a sidecar is the motorcycle 30 on the right side and the sidecar 31 on the left side, but the sidecar 31 may be attached to the right side of the motorcycle 30.

  Moreover, in the said 2nd Embodiment, in addition to suppression of an oversteer state or an understeer state, it suppressed that the vehicle became unstable by the lifting of a wheel. However, the instability of the vehicle due to the lifting of the wheels may occur regardless of the oversteer state or the understeer state. Therefore, it is preferable to suppress the vehicle from becoming unstable due to the lifting of the wheels, separately from the suppression of the oversteer state and the understeer state.

  Further, in the calculation of the target yaw rate Yr shown in FIG. 5, the correction torque Ts is calculated based on the engine torque Eg and the running resistance torque Td. When the vehicle is braking, the running resistance torque Td and the engine torque are calculated. In addition to Eg, the correction torque Ts may be calculated from the braking torque. In this case, since braking is being performed, the engine is in a brake state, and the engine torque Eg is a negative value. Further, the braking torque can be estimated from the control hydraulic pressure (pressurization amount) during the pressurization control using the vehicle brake control device 1 and the M / C pressure sensor value if the brake operation is performed by the driver. .

  The steps shown in each figure correspond to means for executing various processes. Specifically, in the brake ECU 4, the part that executes the process of step 105 is a torque acquisition means, a steering amount acquisition means, an actual turning state acquisition means, the part that executes the process of step 110 is an acceleration / deceleration detection means, and step 120. The part that executes the process corresponds to the change amount acquisition means, the part that executes the process of step 125 corresponds to the reference turning state estimation means, and the part that executes the processes of steps 135 and 145 corresponds to the vehicle stabilization control means. Further, the part that executes the processing of step 215 corresponds to a corrected steering amount estimating means.

  DESCRIPTION OF SYMBOLS 1 ... Brake fluid pressure control device, 2 ... Brake fluid pressure control actuator, 4 Brake ECU, 5 ... Wheel speed sensor, 6 ... Yaw rate sensor, 7 ... Rudder angle sensor, 8 Engine ECU, ... 30 ... Bike, ... 31 ... sidecar

Claims (7)

The vehicle body (30) has a driven wheel (FW) and a drive wheel (RW) in the front-rear direction, and the vehicle body (30) is arranged on a line connecting the driven wheel (FW) and the drive wheel (RW). A vehicle stability control device applied to a vehicle having a side wheel (SW) at a position having a deviation in one of the lateral directions,
Torque acquisition means (105) for acquiring drive torque (Eg) applied to the vehicle;
Steering amount acquisition means (105) for acquiring the steering amount of the vehicle;
An actual turning state obtaining means (105) for obtaining an actual turning state which is an actual turning state of the vehicle;
Acceleration / deceleration detecting means (110) for detecting acceleration / deceleration of the vehicle;
Change amount acquisition means (120) for acquiring a change in the corrected steering amount with respect to a change in the drive torque based on a change in the drive torque and a change in the steering amount when the acceleration / deceleration of the vehicle is detected during the straight traveling of the vehicle. When,
Reference turning state estimating means (125) for estimating a reference turning state of the vehicle;
A vehicle that controls acceleration / deceleration of the vehicle by controlling braking means (1) that generates a braking force on at least one of the driven wheel and the driving wheel based on the actual turning state and the reference turning state. Stabilization control means (135, 145),
The reference turning state estimating means (125) has a corrected steering amount estimating means (215) for estimating a corrected steering amount by the driver, and the steering amount acquired by the steering amount acquiring means (105) and the corrected steering amount estimation. Estimating the reference turning state based on the corrected steering amount estimated by the means (215);
The vehicle stabilization control means (135, 145) includes the actual turning state acquired by the actual turning state acquisition means (105), the reference turning state, and the drive torque acquired by the change amount means (120). A vehicle stabilization control device that controls acceleration / deceleration of the vehicle based on a change in a corrected steering amount with respect to the change.   The vehicle stabilization control means (135, 145) is when the vehicle is turning toward the side wheel (SW) and is in an oversteer state in which the turning degree in the actual turning state is large with respect to the reference turning state. 2. The vehicle stabilization control device according to claim 1, wherein the acceleration / deceleration is controlled so that the acceleration of the vehicle decreases or the deceleration increases. Lateral acceleration acquisition means for acquiring lateral acceleration applied to the vehicle;
The vehicle stabilization control means (135, 145) is a threshold which is a condition under which the lateral acceleration acquired by the lateral acceleration acquisition means is lifted while the vehicle is turning toward the side wheel (SW). 3. The vehicle stabilization control device according to claim 1, wherein the acceleration / deceleration is controlled so that the acceleration of the vehicle decreases or the deceleration increases.   The vehicle stabilization control means (135, 145) is when the vehicle is turning to the side opposite to the side wheels and is in an understeer state where the turning degree of the actual turning state is small with respect to the reference turning state. 4. The vehicle stabilization control device according to claim 1, wherein acceleration / deceleration is controlled such that acceleration of the vehicle decreases or deceleration increases. 5.   The vehicle stabilization control means (135, 145) controls the deceleration of the vehicle to be within a predetermined limit deceleration while the vehicle is turning to the side opposite to the side wheel. The vehicle stabilization control device according to any one of claims 1 to 4. Lateral acceleration acquisition means for acquiring lateral acceleration applied to the vehicle;
The vehicle stabilization control means (135, 145) sets the limit deceleration smaller as the lateral acceleration acquired by the lateral acceleration acquisition means is larger while the vehicle is turning toward the side wheel. The vehicle stabilization control device according to claim 5.   7. The vehicle stabilization control device according to claim 1, wherein the side wheel is a driven wheel, and the side wheel is not provided with a brake mechanism.

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