JPH0635259B2 - Vehicle drive system clutch control device - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION (Industrial field of application) The present invention is used for a transfer device of a four-wheel drive vehicle, a differential device of an automobile, and the like, and changes driving force distribution to front and rear wheels or left and right wheels. The present invention relates to a vehicle drive system clutch control device that controls the clutch engagement force of a drive system clutch.

(Prior Art) As a conventional limited slip differential clutch control device for a vehicle, for example, Japanese Patent Laid-Open No. 61-67629 (Japanese Patent Application No. 59-187).
780) and Japanese Patent Laid-Open No. 61-102321 (Japanese Patent Application No.
9-223487) are known.

The former conventional device is a device of the control content that increases the clutch engagement force as the vehicle speed becomes higher vehicle speed, and the latter conventional device is the input information of the acceleration state,
It was a device of the control content that increases the clutch engagement force as the acceleration increases.

(Problems to be Solved by the Invention) However, in such a conventional device, since the clutch engagement force is controlled only by the input information of the vehicle speed or the acceleration state, the clutch engagement force is generally reduced. When driving on a low friction coefficient road with a device controlled based on a strong high friction coefficient road command value map, the tail swing occurs due to the sudden increase in the differential limiting torque, and the high friction coefficient road is also generated. There was a problem that strong understeer occurred at the turn-in at a tight corner even when driving in the car. On the other hand, when traveling on a high friction coefficient road with a device controlled based on a low friction coefficient road command value map in which the clutch engagement force is weak overall, the wheel spin of the turning inner wheel is low because the differential limiting torque is low. Is noticeable, and there is a problem in that the turning acceleration is poor.

Therefore, it is possible to correct the command value by the road surface friction coefficient, but it is difficult to directly and accurately detect the road surface friction coefficient.

(Means for Solving Problems) The present invention has been made for the purpose of solving the above-mentioned problems, and in order to achieve this object, the present invention provides the following solving means. .

The solution means of the present invention will be described with reference to the conceptual diagram of claims shown in FIG. 1. A power split device 1 for distributing and transmitting an engine driving force to front and rear or left and right drive wheels, and the power split device 1
Drive system clutch means 2 which is provided between the drive input section and the drive output section of the vehicle and generates a transmission torque by an external control force.
A vehicle drive system clutch control device comprising: a detection means 3 for detecting a vehicle state; and a clutch control means 4 for outputting a control signal for increasing or decreasing a clutch engagement force based on an input signal from the detection means 3. The detection means 3 includes a vehicle speed detection means 301, an acceleration operation detection means 302, and a centripetal acceleration operation detection means 303, and the clutch control means 4 increases the vehicle speed or a rapid acceleration state by the clutch. It is characterized in that it is a means for outputting a control signal to increase the engaging force of the means 2 and correcting it so as to increase the engaging force of the clutch means 2 as the centripetal acceleration increases.

(Operation) In the vehicle drive system clutch control device of the present invention, the clutch engagement force corresponding to the vehicle speed and the acceleration state is automatically obtained by the above-mentioned means, and the clutch engagement force increases in accordance with the centripetal acceleration. Will be corrected to.

Therefore, even if the vehicle speed is the same and the turning radius is the same, the centripetal acceleration is large on the high friction coefficient road and small on the low friction coefficient road, and the centripetal acceleration becomes indirect input information of the road friction coefficient, For example, based on clutch engagement force control on low friction coefficient roads, the optimum clutch engagement force for various driving patterns such as turning acceleration running, turning initial running, high speed straight running and low friction coefficient running Control can be performed.

(Examples) Hereinafter, examples of the present invention will be described in detail with reference to the drawings. In describing this embodiment, an automobile differential limiting control device equipped with a multi-plate friction clutch means that is operated by an external hydraulic pressure will be taken as an example.

First, the configuration of the embodiment will be described.

As shown in FIGS. 2 to 4, the embodiment apparatus includes a differential device (power split device) 10, a multi-plate friction clutch means (drive system clutch means) 11, a hydraulic pressure generator 12, a control unit (clutch control). Means) 13 and an input sensor 14, each of which will be described below.

The differential device 10 has a differential function of providing a speed difference between the left and right wheels in accordance with the difference in rotation speed in a traveling state in which a difference in rotation speed occurs between the left and right wheels, and the engine driving force is applied to the left and right drive wheels. It is a device having a driving force distribution function that distributes and transmits the distribution.

The differential device 10 is housed in a housing 16 attached to a vehicle body by a stud bolt 15, and includes a ring gear 17, a differential case 18,
A pinion mate shaft 19, a differential pinion 20, and side gears 21 and 21 'are provided.

The differential case 18 includes a housing 16
On the other hand, it is rotatably supported by tapered roller bearings 22 and 22 '.

The ring gear 17 is a differential case 18
The drive pinion 24 is fixed to the drive pinion 24 and is engaged with the drive pinion 24 provided on the propeller shaft 23, and the rotational drive force is input from the drive pinion 24.

The left gear side drive shaft 25 and the right wheel side drive shaft 2 which are drive output shafts are provided on the side gears 21 and 21 '.
6 are provided for each.

The multi-disc friction clutch means 11 is provided between the drive input portion and the drive output portion of the differential device 10 and is a means for applying a clutch engaging force by an external hydraulic pressure to generate a differential limiting torque.

The multi-plate friction clutch means 11 is housed in the housing 16 and the differential case 18, and comprises multi-plate friction clutches 27, 27 ', pressure rings 28, 28', reaction plates 29, 29 ', thrust bearing 30. , 30 ', spacers 31, 31', push rod 32, hydraulic piston 33, oil chamber 34, and hydraulic port 35.

The multi-plate friction clutches 27, 27 'are fixed in the rotational direction by friction plates 27a, 27'a fixed in the differential case (drive input section) 18 in the rotational direction, and by side gears (drive output sections) 21, 21'. The friction disks 27b and 27b ', and the pressure rings 28 and 28' are formed on both end surfaces in the axial direction.
And reaction plates 29, 29 'are arranged.

The pressure rings 28, 28 'are provided in a fitted state on the pinion mate shaft 19 as a member for receiving a clutch fastening force, and the fitting portion has a rectangular cross section as shown in FIG. The end portion 19a is fitted with the square grooves 28a and 28'a, and the structure is such that, unlike the torque proportional differential limiting means, thrust force due to the left / right wheel rotation difference is not generated.

The hydraulic piston 33 moves in the axial direction (to the right in the drawing) by the hydraulic pressure supplied to the hydraulic port 35 to engage both multi-plate friction clutches 27, 27 'in accordance with the hydraulic level. In the plate friction clutch 27, the fastening force is transmitted to the push rod 32 → the spacer 31 → the thrust bearing 30 → the reaction plate 29, and the pressure ring 2 is transmitted.
8 is fastened as a reaction force receiver, and the other multi-plate friction clutch 27 'is fastened by the fastening reaction force from the housing 16 as a fastening force.

As shown in FIG. 4, the hydraulic pressure generator 12 is an external device that generates a hydraulic pressure that is a clutch engagement force, and is a hydraulic pump 4
0, a pump motor 41, a pump pressure oil passage 42, a drain oil passage 43, a control pressure oil passage 44, and a solenoid proportional pressure reducing valve 4 having a valve solenoid 45 as a valve actuator.
6 is provided.

The electromagnetic proportional pressure-reducing valve 46 supplies the hydraulic oil having a pump pressure supplied from the hydraulic pump 40 via the pump pressure oil passage 42 to a large command current value i ^* according to the control current signal (i) from the control unit 13. The control current signal (i) is an electromagnetic proportional pressure reducing valve 46, which is a valve actuator that controls the control hydraulic pressure P proportional to the pressure and sends the control hydraulic pressure P from the control pressure oil passage 44 to the hydraulic port 35 and the oil chamber 34. Is output to the valve solenoid 45.

The control oil pressure P and the differential limiting torque T are in the relationship of T∝P · μ · n · r · A n; number of clutches r; average clutch radius A; pressure receiving area, and the differential limiting torque T is controlled. Proportional to hydraulic pressure P.

The control unit 13 is a control circuit centered on a vehicle-mounted microcomputer, and includes an input interface circuit 131, a memory 132, a CPU (central processing unit) 133, and an output interface circuit 13.
It is equipped with 4.

In addition, the input sensor 14 to the control unit 13
The left front wheel speed sensor 141 and the right front wheel speed sensor 14
2. An accelerator opening sensor 143 is provided.

The left and right front wheel speed sensors 141 and 142 is a sensor that outputs the left and right front wheels rotational speed _W L, _W left wheel rotational speed signals corresponding to _{R (w L)} and the right front wheel rotational speed signal _{(w R).}

The left and right front wheel speed sensors 141, 142 and the vehicle speed calculating circuit of the control unit 13 constitute vehicle speed detecting means, and the centripetal acceleration calculating circuit of the left and right front wheel speed sensors 141, 142 and control unit 13 form the centripetal acceleration detecting means. Composed.

The accelerator opening sensor 143 is a sensor that detects the degree of depression on the accelerator pedal and outputs an accelerator opening signal (a) corresponding to the accelerator opening A (also referred to as throttle opening).

The accelerator opening sensor 143 and the accelerator opening time change rate calculation circuit of the control unit 13 constitute acceleration operation detecting means.

Next, the operation of the embodiment will be described.

First, the flow of the differential limiting control operation performed by the control unit 13 at a predetermined control cycle will be described with reference to the flowchart shown in FIG.

In step 100, the left and right front wheel speed sensors 141, 142.
And a front left wheel speed W _L , a right front wheel speed W _R, and an accelerator opening A from the accelerator opening sensor 143.

Step 101 is a calculation step for calculating the vehicle speed V from the left front wheel speed W _L and the right front wheel speed W _R.

The vehicle speed V is calculated by the following equation.

However, γ is the tire radius. Step 102 is a calculation step for calculating the centripetal acceleration Y _G from the left front wheel speed W _L , the right front wheel speed W _R, and the vehicle speed V.

The centripetal acceleration Y _G is calculated based on an arithmetic expression according to the theory described below.

Generally, the centripetal acceleration Y _G is: Y _G = R _g · ω ² ... R _g ; Turning radius ω; Turning angular velocity Further, the relationship between the vehicle speed V and the turning angular velocity ω is V = R _g · ω ... On the other hand, the difference between the left and right rotational speeds of the driven wheels ΔW (= | W _L −W
_R |) is proportional to the vehicle speed V and inversely proportional to the turning radius R _g , K ₂ is represented by a proportional constant.

Furthermore, if we transform the equation, Becomes

Therefore, when substituting the ‘expression into the expression, Is obtained.

The centripetal acceleration Y _G obtained by the above analysis is shown in the graph as shown in FIG. 6, where Y _G = K · V
The centripetal acceleration Y _G is calculated by the control unit 13 using the relationship of ΔW.

Step 103 is a calculation step for calculating the accelerator opening time change rate Å based on the current accelerator opening A, the accelerator opening A ₀ read one cycle before, and the control cycle ΔT.

The accelerator opening time change rate Å is calculated by the following equation.

Å = (A−A ₀ ) / ΔT Step 104 is the above-mentioned steps 100, 101, 1
Fig. 5 (a) (b) (c) (d) based on the accelerator opening A, the vehicle speed V, and the accelerator opening time change rate Å obtained in 03.
The provisional command current value i ^** is obtained from the control characteristic map shown in FIG.

The control characteristic map is a map in which the command current value is set based on the low friction coefficient road, and the value of the accelerator opening time change rate Å displayed on the control characteristic map is from zero accelerator opening. The time required for full opening is shown. The larger the value, the slower the acceleration operation, and the smaller the value, the faster the acceleration operation.

Step 105 is a correction calculation step in which the temporary command current value i ^** obtained in step 104 is corrected by the centripetal acceleration Y _G obtained in step 102 to obtain the command current value i ^* .

The following equation is used as the equation for calculating the command current value i ^* .

i ^* = The _{^{i ** + f 3 (Y G}} ) ^{_{Specifically, Y G ≧ Y 2; i}} * = i ** +1 Y G <Y 2; i * = i ** However, _{Y 2} is set centripetal acceleration Alternatively, i ^* = i ^** + K _g × Y _G.

That is, when the centripetal acceleration Y _G is large, the command current value i ^* is increased and the clutch engagement force is strengthened.

Step 106 is an output step of outputting the command current value i ^* finally obtained by the correction in step 105 to the electromagnetic proportional pressure reducing valve 46.

Next, the action of each traveling pattern of the vehicle using the embodiment apparatus will be described with reference to FIG.

Generally, what is desired for a limited slip differential control device is to improve power performance such as acceleration. Specifically, it is to improve the performance of the turning acceleration running and the high-speed straight running shown in FIG. 8. If the clutch engagement force is set to be stronger with the conventional limited slip differential control device, the turning initial running and Performance of low friction coefficient road driving deteriorates.

That is, a large differential limiting torque is required at the time of turning with high centripetal acceleration that can occur due to a high friction coefficient road, and even with a high friction coefficient road, at low centripetal acceleration as in the initial running of the turning, or FIG. It is desirable that the differential limiting torque be small when each of the state quantities shown in (3) is judged to be traveling on a low friction coefficient road.

Therefore, by giving a small differential limiting torque to the control characteristic map that is usually set with the low friction coefficient road as a reference, it is possible to prevent early tail swing on low friction coefficient road running and to prevent understeer during the initial turning. Is also suppressed.

In addition, during high-speed straight running that does not require a particularly large differential limiting torque, the high-speed straight running stability is ensured by slightly increasing the differential limiting torque according to the vehicle speed V in the map set for the low friction coefficient road. To be done.

On the other hand, during turning on a high friction coefficient road, the centripetal acceleration Y _G also becomes a large value and the command current value i ^* also becomes a large value by the correction, so that a large differential limiting torque is obtained.

Therefore, when turning on a high friction coefficient road, the inner ring spin is suppressed by the large differential limiting torque, a large amount of torque is transmitted to the outer wheel, acceleration can be improved, and appropriate turning performance is achieved. become.

This can be realized because the tire has a large capacity (friction force) due to the high friction coefficient road.

By the way, since the tire capacity is small on low friction coefficient roads, if differential torque is limited and a large amount of torque is applied to the outer wheels, the tires spin on both the inner and outer wheels, the lateral force decreases and the entire vehicle spins. Will end up.

Although the embodiment of the present invention has been described in detail above with reference to the drawings, the specific configuration is not limited to this embodiment, and the present invention can be applied even if there is a design change or the like within a range not departing from the gist of the present invention. included.

For example, in the embodiment, the acceleration Y _G is the set centripetal acceleration Y ₂
In the above cases, the clutch engagement force was corrected to be increased, but it is also possible to set the hysteresis characteristic by making the value of the set centripetal acceleration different at the start of correction and at the time of cancellation of correction. is there.

The above control example will be described with reference to FIG. In the figure, the vertical axis represents centripetal acceleration and fastening force correction signals, and the horizontal axis represents time. As is apparent from the figure, the operating side centripetal acceleration set value Y ₂ which is the reference value when starting the clutch engagement force correction is the release side centripetal acceleration set value Y which is the reference value when the engagement force correction is released. It is set as a value larger than ' ₂ . By setting the hysteresis characteristics by making the two set values Y ₂ and Y ′ ₂ different from each other, the centripetal acceleration is caused in a short time due to the steering wheel correction at the time of turning and the change of the vehicle speed. Even if the set value Y ₂ is changed up and down, the clutch engagement force is prevented from changing ON-OFF, a good driving feeling can be secured, and the durability of the clutch facing accompanying the change of the clutch operating state can be secured. It is also possible to prevent deterioration of sex.

In addition, the calculated value of the centripetal acceleration decreases and the set centripetal acceleration Y
Even if it becomes smaller than _{2, the} correction is not canceled for a certain period of time, so that it is possible to continue to apply a stable clutch engagement force even when the detected value of the centripetal acceleration temporarily decreases. Can also be

This control example will be described with reference to FIG. In the figure, the vertical axis indicates the centripetal acceleration, the output state of the correction signal, the output state of the correction release prohibition signal, and the clutch engagement force, and the horizontal axis indicates the time. First, a case will be described as an example where a countersteer operation (steering the front wheels in a direction opposite to the turning direction of the vehicle body) is performed while the vehicle is turning right. When the vehicle is turning to the right by normal steering, the centripetal acceleration is calculated from the vehicle speed obtained from the rotational speed detection value of the front wheels and the rotational speed difference as described above. At that time, when the calculated value of the centripetal acceleration is equal to or greater than the set centripetal acceleration Y ₂ , the clutch fastening force correction signal is output as “1” in order to increase the clutch fastening force as in the above-described embodiment. Next, in order to prevent the vehicle from spinning, when the driver performs countersteering operation, the front wheels are steered to the left from the right steering state through the neutral point even though the vehicle body is turning to the right. . Then, the rotation speed of the front wheels was initially
Although the left wheel speed of the turning outer wheel is faster than the right wheel speed of the turning inner wheel, the relationship is reversed by the countersteering operation. Therefore, the centripetal acceleration calculated from the rotational speed difference of the front wheels and the vehicle speed is calculated as a value smaller than the centripetal acceleration Y ₂ set for T ₁ time in the process of steering the front wheels from right to left by the counter-steering operation. .. However, in the present control, even if the calculated value of the centripetal acceleration decreases and becomes smaller than the set centripetal acceleration Y ₂ , the correction cancellation prohibition signal for prohibiting the clutch engagement force correction cancellation for a certain time (T ₀ ) is set to “1”. Since it is configured to output, even if the calculated value of the centripetal acceleration is temporarily reduced and the correction signal becomes "0", the correction control of the clutch engagement force is continued and the constant clutch engagement force is maintained as shown in the figure.

Thus, since the calculated value of the centripetal acceleration is configured such that outputs a correction canceling inhibit signal to maintain the engagement force correction control for a certain time the clutch even when it becomes smaller than the Y ₂ setting centripetal acceleration decreases It is possible to prevent a reduction in maneuverability due to a change in the differential limiting amount during unstable traveling accompanied by skids that require countersteering operation.

Next, an explanation will be given when the S-shaped corner is normally steered.
When traveling in the S-shaped corner where the vehicle makes a right turn to a left turn, the vehicle goes straight on the way. Therefore, the centripetal acceleration becomes a value smaller than the set centripetal acceleration Y ₂ during the time T ₂ near the straight traveling state, and the correction signal for increasing the clutch engaging force is “0”. In this case as well, similarly, the cancellation of the correction is prohibited during the initial T ₀ time, but the correction is canceled thereafter and the clutch engaging force decreases, so that the turning characteristic does not become a strong oversteer.

Further, in the embodiment, the control example of the differential limiting clutch for limiting the differential between the left and right driving wheels has been shown.
The present invention can be applied to a clutch engagement force control device for a transfer clutch that distributes the drive force to the front and rear wheels of a four-wheel drive vehicle as shown in No. 76048. In this case as well, the clutch engagement force can be increased during turning acceleration traveling. The stability of the vehicle behavior can be ensured by adopting the four-wheel drive side.

Further, in the embodiment, the electromagnetic proportional pressure reducing valve is shown as the actuator, but an example in which an opening / closing electromagnetic valve or the like is used and hydraulic pressure control is performed by using a control signal as a duty signal, an electromagnetic clutch or the like is also used. It may be an example in which a variable differential limiting torque is obtained by the other differential limiting means.

(Effects of the Invention) As described above, in the vehicle drive system clutch control device of the present invention, the clutch control means engages the clutch means as the vehicle speed increases or in the rapid acceleration state. Since the control signal is output to increase the force and the correction is performed so as to increase the engaging force of the clutch means as the centripetal acceleration increases, the road surface friction coefficient and various kinds of friction can be directly detected without directly detecting the road surface friction coefficient. It is possible to obtain the effect that the optimum clutch control according to the traveling pattern can be achieved with a simple control device.

[Brief description of drawings]

FIG. 1 is a conceptual view of a claim showing a vehicle drive system clutch control device of the present invention, FIG. 2 is a sectional view showing a differential device incorporating a differential limiting means of an embodiment of the present invention, and FIG. FIG. 2 is a view in the Z direction, FIG. 4 is a view showing a hydraulic pressure generating device and a control device of the embodiment device, and FIGS. 5 (a), (b), (c) and (d) are set in the control unit of the embodiment. FIG. 6 is a control characteristic map diagram, FIG. 6 is a characteristic diagram showing the relationship between centripetal acceleration with respect to the rotational speed difference and the vehicle speed, FIG. 7 is a flowchart diagram showing the flow of the differential limiting control operation of the embodiment apparatus, and FIG. The figure showing the table showing the state quantity in the pattern, and FIGS. 9 and 10 are control time charts in other control examples. 1 ... Power split device 2 ... Drive system clutch means 3 ... Detection means 301 ... Vehicle speed detection means 302 ... Acceleration operation detection means 303 ... Centripetal acceleration detection means 4 ... Clutch control means

Claims (1)

[Claims] 1. A power split device for distributing and transmitting an engine driving force to front and rear or left and right drive wheels, and a transmission torque generated by a control external force provided between a drive input part and a drive output part of the power split device. Vehicle drive system clutch control including drive system clutch means for detecting the vehicle state, detection means for detecting a vehicle state, and clutch control means for outputting a control signal for increasing or decreasing the clutch engagement force based on an input signal from the detection means. In the device, the detection means includes a vehicle speed detection means, an acceleration operation detection means, and a centripetal acceleration detection means, the clutch control means,
A control signal is output to increase the engagement force of the clutch means as the vehicle speed increases or the acceleration state increases, and correction is performed so that the engagement force of the clutch means increases as the centripetal acceleration increases. A drive system clutch control device for a vehicle, which is a means.