Method for controlling a road vehicle with variable stiffness and rear steering wheels while traveling around a curve - Patents.com

Abstract

To provide a method of controlling, while driving along a curve, a road vehicle with a variable stiffness and with rear steering wheels.SOLUTION: The method comprises the steps of: determining an actual attitude angle (β) of the road vehicle; establishing a desired attitude angle (βTGT); determining an actual yaw rate (dψ / dt) of the road vehicle; establishing a desired yaw rate (dψ / dtTGT); and changing, in a simultaneous and coordinated manner, the steering angle (α) of the rear wheels and the distribution of the stiffness of the connection of the four wheels to the frame depending on a difference (εβ) between the actual attitude angle (β) and the desired attitude angle (βTGT) and depending on a difference (εψ) between the actual yaw rate (dψ / dt) and the desired yaw rate (dψ / dtTGT).SELECTED DRAWING: Figure 5

Description

[Technical field]

[0001] [CROSS REFERENCE TO RELATED APPLICATIONS] This patent application claims priority to Italian Patent Application No. 102020000002746, filed February 12, 2020, the disclosure of which is incorporated herein by reference in its entirety.

[0002] The present invention relates to a method for controlling a road vehicle with variable roll stiffness and rear steerable wheels while traveling around a curve. [Background technology]

[0003] Four-wheel steering vehicles, also known as 4WS vehicles, have been produced by the automotive industry for many years, but there are few models available and they are primarily based on passive mechanical systems that can only provide a "fixed" steering angle to the rear wheels. In these passive mechanical systems, the rear wheels are coupled to the front wheels in a way that provides a fixed steering angle to the rear wheels when the front wheels are steered.

[0004] More modern dynamic systems, comprising active actuators (electrical or hydraulic) controlled by an electronic control unit, make it possible to control the steering of the rear wheels by imparting to them a variable steering angle within a given range.

[0005] While driving at high speeds (e.g., above 60 km / h) along a curve, the steering of the rear wheels is performed to increase the stability of the vehicle, especially when it is close to grip loss, so that the rear wheels are steered in phase (i.e., steered in the same direction as the steering of the front wheels, i.e., when the front wheels steer to the right, the rear wheels also steer to the right) to reduce the yaw angle and thereby make the vehicle easier to control. In other words, the steering of the rear wheels in phase counteracts the centrifugal force that tends to make the rear of the vehicle skid while driving along a curve, thereby allowing the rear of the vehicle to conform to an ideal trajectory, improving stability and effectiveness. In this situation, the rear wheels are usually steered at a steering angle that is smaller than 1-2° and depends on the steering angle of the front wheels.

[0006] When steering at low speeds (e.g. up to 60 km / h), the steering radius of the rear wheels is significantly reduced and the rear wheels are steered in anti-phase (i.e. in the opposite direction to the steering of the front wheels, i.e. when the front wheels steer to the right the rear wheels steer to the left and vice versa) with a maximum steering angle of 3-5° depending on the steering angle of the front wheels in order to increase the maneuverability of the vehicle especially at low speeds.

[0007] US Patent No. 5,399,433 describes a method for controlling a road vehicle with rear steered wheels while traveling along a curve, the control method comprising the steps of determining an actual attitude angle of the road vehicle, determining a desired attitude angle, and varying the steering angle of the rear wheels depending on the difference between the actual attitude angle and the desired attitude angle.

[0008] 3, No. 5, pp. 1111-1115, 2003 describes a method for controlling a road vehicle with variable stiffness and rear steering wheels while traveling along a curve. [Prior art documents] [Patent documents]

[0009] [Patent Document 1] European Patent Application Publication No. 3153382 [Non-patent literature]

[0010] [Non-Patent Document 1] Liang Wu et al., "The Integrated Effects of Active Suspension and Rear Steering Control System on Vehicle Lateral Stability", Konferenzbeitrag“IBEC 2003” und SAE Technical Paper Series 2002-01-20 48, vol.1, 28 March 2017, XP055738637, US ISSN:0148-7191, DOI:10.4271 / 2017-01-0257 Summary of the Invention [Problem to be solved by the invention]

[0011] It is an object of the present invention to provide a method of controlling a road vehicle with variable stiffness and rear steered wheels while traveling around a curve, said control method maximizing the performance of the road vehicle while traveling around a curve without destabilizing the road vehicle, while at the same time being simple and economical to implement. [Means for solving the problem]

[0012] According to the present invention there is provided a method of controlling a road vehicle with variable roll stiffness and rear steerable wheels while navigating a curve, as claimed in the accompanying claims.

[0013] The appended claims describe preferred embodiments of the invention and form an integral part of the body of the specification. [Brief description of the drawings]

[0014] The present invention will now be described with reference to the accompanying drawings, which show non-limiting embodiments of the invention. [Figure 1] 1 is a schematic plan view of a road vehicle with rear steered wheels controlled in accordance with the present invention; [Diagram 2]FIG. 2 is a schematic plan view of the road vehicle of FIG. 1 highlighting the rear wheel steering mechanism. [Diagram 3] FIG. 2 is a schematic plan view of the road vehicle of FIG. 1 while traveling around a curve, highlighting the trajectory, driving speed, and attitude angle. [Figure 4] FIG. 2 is a schematic perspective view of the road vehicle of FIG. 1 with emphasis on the yaw angle. [Diagram 5] FIG. 2 is a block diagram showing the logic of the control method according to the present invention. DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

[0015] In FIG. 1, the numeral 1 generally indicates a road vehicle with two front wheels 2 and two rear wheels 3 , which receive a torque from a powertrain system 4 .

[0016] The powertrain system 4 comprises an internal combustion engine 5 arranged in a front longitudinal position and having a crankshaft 6, and a servo-assisted drivetrain 7 which transmits the torque generated by the internal combustion engine 5 to the rear drive wheels 3 and has a configuration called "transaxle". The drivetrain 7 comprises a driveshaft 8 connected on one side to the crankshaft 6 and on the other side mechanically connected to a transmission 9 which comprises at least one clutch and is arranged in a rear longitudinal position. The transmission 9 is connected in a train-like manner to an electronically controlled self-locking differential 10 from which a pair of axle shafts 11 start, each axle shaft being integral with a respective rear drive wheel 3.

[0017] Each wheel 2, 3 is mechanically connected to the frame 12 of the road vehicle 1 by a suspension 13 (partially shown in FIG. 1 ) that is equipped with an electronically controlled actuator 14 that makes it possible to vary (increase or decrease) both the elastic stiffness and the damping of the suspension elements. By way of example, the actuator 14 can adjust both the set vertical stiffness and the damping constants of the suspension 13 in parallel with the conventional elastic elements of the suspension 13 (helical springs and passive hydraulic shock absorbers). By way of example, this adjustment can be performed by different elements such as electronically controlled anti-roll bars and magnetorheological shock absorbers.

[0018] The electronically controlled actuators 14 make it possible to vary the stiffness of the connection of the four wheels 2, 3 to the frame 12 of the road vehicle 1 and thus to modify the distribution of the vertical loads acting on the wheels 2, 3 over the four wheels 2, 3 in dynamic situations, i.e. when there are longitudinal or lateral accelerations. In other words, by varying the stiffness of the electronically controlled actuators 14, it is possible to vary the distribution of the vertical loads over the four wheels 2, 3 and thus the punctual vertical load acting on each wheel 2, 3.

[0019] According to FIG. 2 , the rear drive wheels 3 are steered wheels, i.e. they are supported by corresponding suspensions 13 so that they can rotate together (i.e. with the same degree of rotation) to the right or left about a vertical axis in order to vary the corresponding steering angle α (shown in FIG. 3 ); in particular, an (electrical or hydraulic) actuator 15 is provided that actively controls the change in the steering angle α of the rear drive wheels 3.

[0020] The steering of the rear drive wheels 3 increases the stability of the road vehicle 1 while driving at high speed (e.g. above 60 km / h) around a curve, and in particular to improve the stability of the road vehicle 1 while the vehicle is driving at high speed around a curve, the rear drive wheels 3 are steered in phase (i.e. steered in the same direction as the steering of the front wheels 2, i.e. when the front wheels 2 steer to the right, the rear wheels 3 also steer to the right). In other words, the steering of the rear drive wheels 3 in phase counteracts the centrifugal forces that tend to make the rear of the vehicle skid while driving around a curve, thereby allowing the rear of the vehicle to conform to an ideal trajectory, improving stability and effectiveness. In this situation, the rear drive wheels 3 are steered with a steering angle α that is usually smaller than 1-2° and usually depends on the steering angle of the front wheels 2 (i.e. the larger the steering angle of the front wheels 2, the larger the steering angle α of the rear drive wheels 3).

[0021] When steering at low speeds (e.g. up to 60 km / h), the rear drive wheels 3 are steered in antiphase (i.e. in the opposite direction to the steering of the front wheels 2, i.e. when the front wheels 2 steer to the right the rear drive wheels 3 steer to the left and vice versa) with a maximum angle of 3-5° depending on the steering angle of the front wheels 2 (i.e. the greater the steering angle of the front wheels 2, the greater the steering angle α of the rear drive wheels 3), in order to significantly reduce the steering radius and increase the maneuverability of the road vehicle 1 especially at low speeds.

[0022] The road vehicle 1 comprises an electronic control unit 16 ("ECU") which regulates the behaviour of the road vehicle 1 whilst it is travelling along a curve by acting on the electronically controlled actuators 14 and 15 of the suspension 13 in order to control, inter alia, the steering angle α of the rear drive wheels 3, as will be explained in more detail below. The control unit 16 may physically consist of one single device or of different separate devices which communicate with each other via the CAN network of the road vehicle 1.

[0023] According to Fig. 3, when traveling along a curve, the control unit 16 determines, in a known manner, the attitude angle β of the road vehicle 1 (i.e. the angle between the longitudinal axis x of the road vehicle 1 and the direction of the travel speed V of the road vehicle 1 at the centre of gravity B). As an example, the control unit 16 estimates the trajectory T followed by the road vehicle 1 using measurements provided in real time by the three-axis gyroscope and by the GPS tracking unit, in particular the trajectory T is determined by integrating twice the accelerations measured by the three-axis gyroscope, and the measurements provided by the GPS tracking unit are used to periodically cancel position errors occurring during the integration process. Furthermore, the control unit 16 estimates the driving speed V of the road vehicle 1 at the centre of gravity B using the measurements provided in real time by the three-axis gyroscope, in particular the speed V of the road vehicle 1 at the centre of gravity B being determined by one integration of the acceleration measured by the three-axis gyroscope (to make sure that the driving speed V of the road vehicle 1 at the centre of gravity B is in fact tangent to the trajectory T followed by the road vehicle 1, otherwise in case of significant deviations at least one further iteration of the calculations is performed with correction of the parameters used).

[0024] While traveling around the curve, the control unit 16 determines in real time (e.g., as described below) the actual (true) attitude angle β of the road vehicle 1. Furthermore, while traveling around the curve, the control unit 16 determines the desired (ideal) attitude angle β that allows maximizing performance (i.e., driving speed around the curve) while keeping the vehicle 1 stable (i.e., completely safe). TGT Determine.

[0025] According to a possible (although not binding) embodiment, the control unit 16 periodically (for example at a frequency of at least a few tens of Hz and in a known manner) estimates the grip of the wheels 2, 3 with respect to the road surface, determines the radius of curvature of the trajectory T of the road vehicle 1 (i.e. determines the degree of curvature of the trajectory T) and determines the driving speed V of the road vehicle 1. Depending on the grip of the wheels 2, 3 (and therefore the stability of the road vehicle 1), the radius of curvature of the trajectory T and the driving speed V, the control unit 16 determines the desired attitude angle β TGT and in particular the desired attitude angle β TGT To determine this, the control unit 16 may use a predefined map (usually established experimentally during the design and setup phase of the road vehicle 1 and stored in memory) or a mathematical model of the road vehicle 1.

[0026] While traveling along the curve, the control unit 16 determines, in a known manner, the actual yaw rate dψ / dt of the road vehicle 1, which is the rate of change of the yaw angle ψ (shown in FIG. 4), i.e. the first derivative of the yaw angle ψ. As is known, the yaw angle ψ describes the oscillation of the road vehicle 1 about a vertical axis Y passing through the center of gravity B of the road vehicle 1.

[0027] While driving along the curve, according to FIG. 5, the control unit 16 determines the actual attitude angle β of the road vehicle 1 and the desired attitude angle β TGT , determine the actual yaw rate dψ / dt of the road vehicle 1, and determine the desired yaw rate dψ / dt TGT and then, in a simultaneous and coordinated manner, the actual attitude angle β and the desired attitude angle β TGT The difference ε between β Depending on the actual yaw rate dψ / dt and the desired yaw rate dψ / dt TGT Depending on the difference εψ between the steering angle α of the rear wheels 3 and the distribution of the stiffness of the connections of the four wheels 2, 3 to the frame 12 are modified.

[0028] FIG. 5 illustrates the above diagrammatically, and shows how the control unit 16 includes a calculation block 17 for determining the actual attitude angle β of the road vehicle 1, a calculation block 18 for determining the actual yaw rate dψ / dt of the road vehicle 1, a calculation block 19 for determining the desired attitude angle β TGT and the desired yaw rate dψ / dt TGT A calculation block 19 for determining the actual attitude angle β and the desired attitude angle β TGT A subtractor block 20 that calculates the difference εβ between the actual yaw rate dψ / dt and the desired yaw rate dψ / dt TGT and finally a calculation block 22 which controls in a simultaneous and coordinated manner the electronically controlled actuators 14 (to vary the distribution of the stiffness of the connections of the four wheels 2, 3 to the frame 12) and the actuators 15 (to vary the steering angle α of the rear wheels 3).

[0029] According to a preferred embodiment, the calculation block 22 calculates together (i.e. in a simultaneous and coordinated manner) the difference between the two ε β and ε ψ In response to this, the desired steering angle α of the rear wheels 3 is TGT (with which the calculation block controls the actuators 15) and the desired distribution D of the stiffness of the connections of the four wheels 2, 3 to the frame 12 TGT (which is used by the calculation block to control the electronically controlled actuator 14).

[0030] According to a preferred embodiment, the calculation block 22 calculates the desired steering angle α of the rear wheels 3 by combining an open loop control logic with a feedback control logic. TGT and the desired distribution D of the stiffness of the connections of the four wheels 2, 3 to the frame 12 TGT Determine.

[0031] Furthermore, according to a preferred embodiment, the desired steering angle α of the rear wheels 3 TGT and the desired distribution D of the stiffness of the connections of the four wheels 2, 3 to the frame 12 TGTare determined by one single common mathematical model (thus jointly adjusting the effects on the steering and on the stiffness distribution of the rear wheels 3).

[0032] According to a preferred embodiment, the actual attitude angle β and the desired attitude angle β TGT are compared in the time domain, i.e., the actual attitude angle β is compared at each instant with the desired attitude angle β TGT On the other hand, the actual yaw rate dψ / dt and the desired yaw rate dψ / dt TGT are compared in the frequency domain. The time domain analysis applied to the actual attitude angle β, on the other hand, shows its description in terms of the frequency range (spectrum) of the yaw rate dψ / dt TGT In contrast to the frequency domain analysis applied to the yaw rate dψ / dt, we present its standard description for the tine variables. TGT is considered as a superposition of complex sine waves (whose respective phases are typically ignored), each representing a given frequency.

[0033] According to a preferred embodiment, the calculation block 22 of the control unit 16 calculates the vehicle speed while driving along a curve under low to medium lateral acceleration conditions (typically lateral accelerations between 4 and 6 m / s 2 Generally, the lateral acceleration is smaller than 4 to 6 m / s 2 Less than a threshold in the range, e.g. 5m / s 2 α of the rear wheels 3 and the desired distribution D of the stiffness of the connections of the four wheels 2, 3 to the frame 12. TGT The adjustment between the steering angle α of the rear wheels 3 and the stiffness distribution D TGT In other words, while driving along a curve under low to medium lateral acceleration conditions, the master action is performed uniquely by the change in the steering angle α of the rear wheels 3, while the desired distribution D of the stiffness of the connections of the four wheels 2, 3 to the frame 12 isTGT is adjusted simply as a slave to changes in the steering angle α of the rear wheels 3 (i.e., adjusted to accommodate changes in the steering angle α of the rear wheels 3).

[0034] According to a preferred embodiment, the calculation block 22 of the control unit 16 is programmed such that operation involves maximizing lateral stability under high lateral acceleration conditions (generally meaning close to the lateral stability limit, i.e. when the lateral acceleration is close to the lateral stability limit). As an example, the lateral acceleration is 1-2 m / s from the maximum lateral acceleration which represents the lateral stability limit. 2 If the vehicle is too far away, it is close to the lateral stability limit, i.e., the difference between the current lateral acceleration and the maximum lateral acceleration that represents the lateral stability limit is 1 to 2 m / s 2 Less than the threshold in the range, for example 1.5m / s 2 From another perspective, the lateral stability limit is close when the current lateral acceleration reaches (exceeds) 85% of the maximum lateral acceleration, which represents the lateral stability limit. It should be pointed out that the maximum lateral acceleration is not a constant value that applies to all conditions, but is periodically updated (estimated) by the control unit 16, and in particular the maximum lateral acceleration varies essentially depending on the grip conditions due to the interaction between the tires and the road surface.

[0035] The steering angle α of the rear wheels 3 and the desired distribution D of the stiffness of the connections of the four wheels 2, 3 to the frame 12 TGT The coordination between the front and rear axle shafts 2, 3 and 4 favours maximising the lateral force generating capacity of the front axle shafts (consisting of the two front wheels 2) and the rear axle shafts (consisting of the two rear wheels 3), which varies depending on the curve covering conditions, i.e. the presence of longitudinal driving or braking forces (at the start or end of the curve) or simply curb covering. In this situation the main role is played by the distribution D of the stiffness of the connections of the four wheels 2, 3 to the frame 12. TGTThis action is then associated with an in-phase steering of the steering angle α of the rear wheels 3. In other words, while driving along a curve under high lateral acceleration conditions, the master action is the desired distribution D of the stiffness of the connections of the four wheels 2, 3 to the frame 12. TGT On the other hand, the steering angle α of the rear wheels 3 is uniquely implemented by the variation of the desired stiffness distribution D TGT (i.e., the desired distribution D TGT (adjusted to fit variations in

[0036] In order to maximize the lateral force generating capacity of the front and rear wheels 2, 3 in this particular dynamic situation, it is now convenient to reduce the height of the centre of gravity of the road vehicle 1 through adjustment of the stiffness of the connections of the four wheels 2, 3 to the frame 12, by selectively acting on the electronically controlled (front and rear) actuators 14 on the inside or outside of the curve.

[0037] According to a preferred embodiment, the calculation block 22 of the control unit 16 is programmed to be able to carry out the adjustments in a coordinated manner, taking into account the dynamic need to shorten the response time of the rear axle by a fast response of the forces generated by the rear wheels 3 following the steering command of the front wheels 2 required by the driver, for example to reduce the roll movements of the road vehicle in the presence of the combined action of steering and vertical vibrations, or to optimize the characteristic angle variations of the suspension 13, while driving along a curve and when the operating conditions are intermediate between the two aforementioned situations (i.e. between low and medium lateral acceleration conditions and high lateral acceleration conditions), and the need to adjust the elastic stiffness. In other words, when the operating conditions are intermediate between the two aforementioned situations (i.e. between low and medium lateral acceleration conditions and high lateral acceleration conditions), the master action is to adjust the desired distribution D of the stiffness of the connections of the four wheels 2, 3 to the frame 12, so that the desired distribution D of the stiffness of the connections of the four wheels 2, 3 to the frame 12 is achieved. TGT and the steering angle α of the rear wheels 3, i.e. the desired distribution D TGTand the variation of the steering angle α of the rear wheels 3 are coordinated at the same hierarchical level (i.e. they are both "master" actions, one of them does not dominate the other).

[0038] In the embodiment shown in the accompanying figures, the control unit 16 acts on the electronically controlled actuators 14 in order to vary the dynamic response of the suspensions 13 connecting the wheels 2, 3 to the frame 12, i.e. the electronically controlled actuators 14 allow a change in the dynamic response of the suspensions 13 connecting the wheels 2, 3 to the frame 12 and thus a change in the distribution of the roll stiffness of the connection of the four wheels 2, 3 to the frame 12.

[0039] As already mentioned above, the electronically controlled actuators 14 may comprise four electronically controlled shock absorbers (one for each wheel 2, 3) and / or two electronically controlled anti-roll bars (front and rear respectively).

[0040] The embodiments described herein can for this reason be combined with one another without exceeding the scope of protection of the invention.

[0041] The above described control method has several advantages.

[0042] First, the above-described control method maximizes the performance of the road vehicle 1 while traveling around a curve (i.e., maximizes the curve-covering speed) without destabilizing the road vehicle 1 (i.e., maintaining an adequate safety margin but approaching the stability limits of the vehicle 1).

[0043] Furthermore, the control method described above is particularly safe, since it keeps the road vehicle 1 under control at all times and always allows fast and effective action to be taken if necessary.

[0044] The above described control method maximizes the dynamic response speed of the vehicle according to the steering command desired by the driver for the front wheels 2.

[0045] Finally, the above control method, in addition to maximizing the performance of the road vehicle as described above, also enables integrated control of the road vehicle 1 in its entire field of use, i.e., from low to high lateral accelerations, by implementing a complex adjustment of the actuators, while meeting the need to adjust the characteristic angles of the suspension 13 and reduce the roll motion of the vehicle body. [Explanation of symbols]

[0046] 1 vehicle 2 Front wheels 3 Rear wheel 4 Powertrain System 5 Engine 6. Crankshaft 7. Drivetrain 8 Drive shaft 9. Mechanical Transmission 10 Self-locking differential 11 Axle shaft 12 Frames 13. Suspension 14 Electronically controlled actuator 15 Rear steering wheel actuator 16 Control Unit 17 Calculation Blocks 18 Calculation Blocks 19 Calculation Blocks 20 Subtractor Block 21 Subtractor Block 22 Calculation Blocks B Center of gravity β attitude angle ψ Yaw angle Z vertical axis Fx longitudinal force Fy lateral force X vertical axis Y Horizontal Axis Z vertical axis α Rear wheel steering angle D Stiffness distribution

Claims

1. A method for controlling a road vehicle (1) with variable stiffness and rear steering wheels (3) when driving along a curve, said road vehicle (1) comprising a frame (12), four wheels (2, 3), at least one first actuator (14) designed to vary the stiffness distribution of the connections of said four wheels (2, 3) to said frame (12) and at least one second actuator (15) designed to vary the steering angle (α) of the rear wheels (3), said method comprising: determining an actual attitude angle (β) of the road vehicle (1); The desired attitude angle (β ＴＧＴ ) determining an actual yaw rate (dψ / dt) of the road vehicle (1); Desired yaw rate (dψ / dt ＴＧＴ ) The actual attitude angle (β) and the desired attitude angle (β ＴＧＴ ) and depending on the difference (εβ) between the actual yaw rate (dψ / dt) and the desired yaw rate (dψ / dt ＴＧＴ varying in a simultaneous and coordinated manner the steering angle (α) of the rear wheels (3) and the distribution of stiffness of the connections of the four wheels (2, 3) to the frame (12) as a function of the difference (εψ) between and While driving along a curve and the lateral acceleration is 4-6m / s ２ when the steering angle of the rear wheels (3) is less than a threshold value in the range of 0.05 to 0.5 mm, the steering angle of the rear wheels (3) is maximized to maximize the dynamic response speed of the road vehicle (1) to a steering command of the front wheels (2) requested by the driver, and adjusting the stiffness distribution of the connections of the four wheels (2, 3) to the frame (12) accordingly by the first actuator (14).

2. While driving along a curve and the lateral acceleration is between 4 and 6 m / s ２ and less than a stability limit, the steering angle (α) of the rear wheel (3) and the distribution of stiffness of the connection of the four wheels (2, 3) to the frame (12) (D ＴＧＴ 2. The method of claim 1, further comprising the further step of simultaneously adjusting the first and second hierarchical levels.

3. 3. The method of claim 1, wherein the lateral acceleration is considered to be close to the lateral stability limit when the current lateral acceleration reaches 85% of a maximum lateral acceleration representative of the lateral stability limit.

4. The actual attitude angle (β) and the desired attitude angle (β ＴＧＴ ) and depending on the difference (εβ) between the actual yaw rate (dψ / dt) and the desired yaw rate (dψ / dt ＴＧＴ ) depending on the difference (εψ) between the desired steering angle (α ＴＧＴ ) and the desired distribution of stiffness of the connections of the four wheels (2, 3) to the frame (12) (D ＴＧＴ 4. A method according to claim 1, further comprising the further step of determining jointly:

5. The desired steering angle (α ＴＧＴ ) and the desired distribution of stiffness of the connections of the four wheels (2, 3) to the frame (12) (D ＴＧＴ 5. The method of claim 4, wherein: ##EQU1## is determined by combining open loop control logic and feedback control logic together.

6. The actual attitude angle (β) and the desired attitude angle (β ＴＧＴ 6. The method of claim 1, wherein the comparison is made in the time domain.

7. The actual yaw rate (dψ / dt) and the desired yaw rate (dψ / dt ＴＧＴ 7. The method of claim 1, wherein the comparison is made in the frequency domain.

8. 8. The control method according to claim 1, wherein the first actuator (14) changes the dynamic response speed of a suspension (13) connecting the wheels (2, 3) to the frame (12), thereby enabling a change in the distribution of the roll stiffness of the connection of the four wheels (2, 3) to the frame (12) and comprising an electronically controlled shock absorber.

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