JP5071724B2 - Vehicle drive control device - Google Patents

Description

  The present invention relates to a drive control device for a vehicle having a CVT.

As a friction transmission type continuously variable transmission (CVT), there are a belt type and a toroidal type. A friction transmission type CVT variator is usually composed of three elements: a pair of external rotating bodies and an intermediate transmission member sandwiched between the external rotating bodies.
In the belt type, the external rotating body is a conical sheave, and the intermediate transmission member is a belt. In the toroidal type, the external rotating body is a pair of disks, and the intermediate transmission member is a roller.

The toroidal CVT has a line connecting the contact points between the two disks and the roller, and a line connecting the contact points between the two disks and the roller. It is roughly divided into a full toroidal type that passes through.
As a full toroidal CVT, the CVT of Patent Document 1 has been proposed.
WO2007 / 040164 A1

Generally, in the friction transmission type CVT, in order to apply a load to the contact point between the external rotating body and the intermediate transmission member, a clamping force for clamping the intermediate transmission member is applied to the external rotating body. The transmission torque of the external rotating body is obtained by multiplying the sum of the traction forces acting on the external rotating body by the contact radius (the distance between the contact point and the rotation center of the external rotating body).
The traction force F _t generated at the contact point is represented by the product of the slip-dependent traction coefficient μ and the contact load F ₁ . That is,
F _t = μ (x, u) · F ₁ (x, u)
The traction force can be determined experimentally. Traction force depends on two variables: slip and contact load. Slip is a function of the speed of the external rotating body (system input u) and the position and speed of the intermediate transmission member (system state x). The contact load is a function of the clamping force of the external rotating body and the driving force (system input u) and position (system state x) of the intermediate transmission member.

  When the input of the system (the driving force of the intermediate transmission member) directly determines the gear ratio, that is, the sensitivity of the geometric gear ratio (corresponding to the ratio of the contact radius on the input side and the output side) against external rotational speed fluctuations The inventor of the present application has defined that the CVT is a gear ratio control when the is low. As the transmission ratio control CVT, there are a belt type CVT and a half toroidal type CVT. The gear ratio control is realized by controlling the clamp hydraulic pressure using the feedback of the gear ratio.

  On the other hand, when the input of the system directly determines the transmission torque (variator input / output torque), that is, when the clamping force and the driving force of the intermediate transmission member are kept constant, The inventor defined that the CVT is torque control when the speed ratio sensitivity is high. As the torque control CVT, there is a full toroidal CVT.

The torque control type CVT has an internal interaction with a strong torque and gear ratio. Therefore, at the stage of designing the torque control type CVT, it is necessary to consider ensuring the stability of the power train under all conditions, but it is very difficult.
In addition, in the torque control type CVT, the sensitivity of the gear ratio with respect to the rotational speed fluctuation from the outside is high, and for this reason, the gear ratio tends to become unstable, for example, the convergence of the gear ratio at the time of gear change becomes worse. As a result, drivability may be reduced.

  An object of the present invention is to provide a vehicle drive control device capable of reducing the sensitivity of a gear ratio with respect to a change in rotational speed from the outside in a torque control type CVT.

In the transmission ratio control type CVT, the transmission ratio and the control input have a direct relationship. Moreover, the sensitivity of the gear ratio change with respect to external speed fluctuation is high.
On the other hand, in the torque control type CVT, the transmission torque and the control input have a direct relationship. Moreover, the sensitivity of the gear ratio change with respect to external speed fluctuation is high.
And this inventor paid attention to continuous conversion of torque control and gear ratio control. Regarding the conversion between the two controls, it is considered that the conversion from the torque control to the gear ratio control is actually more robust than the reverse conversion. This is because the internal dynamics of the former conversion are considered to be more flexible with respect to control system conversion than the internal dynamics of the latter conversion.

Therefore, the present invention provides an input member (13, 14), an output member (15, 16), and an intermediate transmission member (17) biased with a predetermined piston differential pressure (p) between the input member and the output member. , 18), and a torque control type continuously variable transmission (5) (hereinafter also referred to as CVT) capable of changing the gear ratio (r _CVT ) in a stepless manner, CVT sensitivity reduction means for reducing the sensitivity of the transmission ratio of; Bei example a (52 61, 62), the CVT includes an input disc (13, 14) as an input member which is biased in a direction to approach each other and the output An output disk (15, 16) as a member, a roller (17, 18) as an intermediate transmission member disposed in a toroidal gap formed between both disks, and transmitting torque between the two disks, To be rotatable A roller hydraulic cylinder (1) and a second hydraulic chamber (22, 23) for generating a differential pressure for applying a pulling force to the disks to the rollers via the carriage (19). 21), a first control valve (28) for generating hydraulic pressure in the first oil chamber, a second control valve (30) for generating hydraulic pressure in the second oil chamber, A first opening degree adjusting valve (61) interposed between one oil chamber and the first control valve, and a second opening degree interposed between the second oil chamber and the second control valve. A regulating valve (62), and a third oil chamber (20) for end load to which a higher one of the hydraulic pressures of the first and second oil chambers is guided to bring the input disk closer to the output disk. The first and second opening degree adjusting valves function as the sensitivity lowering means by reducing the opening degree. Then, the controller controls the hydraulic pressure of the third oil chamber based on the actual CVT input torque so as to suppress the slip of the roller with either one of the first and second opening degree adjusting valves closed. It has a function to control the above. In the present invention, since the sensitivity of the CVT gear ratio with respect to fluctuations in the rotational speed from the outside can be reduced, the gear ratio can be stabilized and drivability can be improved.

Moreover , the sensitivity of the transmission gear ratio with respect to fluctuations in the rotational speed from the outside can be reduced in hardware by reducing the opening of the first and second opening adjustment valves. The opening control valve may be a variable throttle or an on / off valve (open / close valve).
Also, a third oil chamber (20 for end load) to which the higher one of the oil pressures of the first and second oil chambers is guided in order to bring the input disk as the input member closer to the output disk as the output member. ), And the control unit closes one of the first and second opening degree adjusting valves and closes the third oil chamber based on the actual CVT input torque so as to suppress the slip of the roller. Therefore , slip can be easily prevented by controlling the pressure of the end load.

The CVT of the torque control type may include a full toroidal type variator (12) (claim 2).

Preferred embodiments of the present invention will be described with reference to the accompanying drawings.
Figure 1 is a schematic diagram showing the drive control apparatus is applied schematic configuration of a drive train 1 of the vehicle one reference embodiment of the present invention. Referring to FIG. 1, drive train 1 includes CVT 5 (hereinafter referred to as CVT 5) formed of a full toroidal continuously variable transmission connected to output shaft 3 of engine 2 via torque converter 4. The CVT 5 includes a CVT input shaft 6 and a CVT output shaft 7.

The CVT output shaft 7 is connected to the drive wheels 11 via a gear train 8, a differential device 9 and a drive shaft 10. The gear train 8 includes a gear 8 a that rotates integrally with the CVT output shaft 7, and a gear 8 b that meshes with the gear 8 a and rotates together with the case of the differential device 9.
The CVT output shaft 7 is formed in a hollow shape, and the CVT input shaft 6 is inserted into the CVT output shaft 7. Input disks 13 and 14 as a pair of input members are provided on the CVT input shaft 6 so as to be integrally rotatable. These input disks 13 and 14 are arranged back to back and form a toroidal lace.

The CVT output shaft 7 is provided with output disks 15 and 16 as a pair of output members, each of which forms an opposing toroidal race on the toroidal race of the pair of input disks 13 and 14, respectively, so as to be integrally rotatable. ing.
Between the toroidal race of the input disk 13 and the output disk 15, a roller 17 for transmitting torque is arranged between the disks 13, 15, and similarly, between the toroidal race of the input disk 14 and the output disk 16, A roller 18 is disposed between the disks 14 and 16 for transmitting torque. Torque is transmitted from the input disk 13 to the output disk 15 via the roller 17, and torque is transmitted from the input disk 14 to the output disk 16 via the roller 18.

The input disks 13 and 14, the output disks 15 and 16 and the rollers 17 and 18 constitute the variator 12 of the CVT 5.
Each roller 17 is supported by a carriage 19 as shown in FIG. 2A. The axis of the carriage 19 extends in a direction orthogonal to the rotation axis of the roller 17 and forms a predetermined caster angle β.

End loads are applied to both the disks 13, 14; 15, 16 by the oil pressure of the oil chamber 20 for end loads (terminal loads). On the other hand, the roller 17, through the carriage 19 receives the biasing force F _P by the pressure difference between the first and second oil chambers 22 of the hydraulic cylinder 21 is pressed against the two discs 13 and 15 . Further, the roller 17 receives a pressing reaction force F _r from each of the disks 13 and 15. Similarly, the roller 18 is pressed against both the disks 14 and 16.

The roller 17 supported by the carriage 19 is arranged around the axis of the carriage 19 in order to eliminate the imbalance between the reaction force generated in the carriage 19 by transmitting torque and the torque necessary to drive the output disk 15. The rotation axis 17a of the roller 17 is inclined so as to produce a swing angle. Thus, as shown in FIG. 2B, the posture is changed in the roller 17, contact with the contact radius R _i (corresponding to the distance from the central axis 6a of the CVT input shaft 6 to the contact point) and the output disc 15 relative to the input disc 13 The radius _Ro changes. The same applies to the roller 18. As a result, the speed ratio between the input disks 13 and 14 and the output disks 15 and 16 changes continuously.

A shuttle valve 25 is interposed in an oil passage 24 connecting the first and second oil chambers 22 and 23 of the hydraulic cylinder 21. Of the first and second oil chambers 22, 23, the oil pressure is supplied from the higher oil chamber to the terminal load oil chamber 20 via the shuttle valve 25 and the oil passage 26. ing.
On the other hand, the first oil chamber 22 of the hydraulic cylinder 21 is supplied with hydraulic pressure from a first hydraulic pump 27 controlled by a first flow control valve 28. In addition, the hydraulic pressure from the second hydraulic pump 29 is supplied to the second oil chamber 23 of the hydraulic cylinder 21 by being controlled by the second flow control valve 30. Instead of the first and second flow control valves 28 and 30, a pressure control valve may be used.

Specifically, the first hydraulic pump 27 is connected to the first hydraulic pump 27 and the first flow rate control valve 28 via a branch point 31a and an oil path 32 in the middle of the oil path 31 connecting the first hydraulic pump 27 and the first flow control valve 28. Oil pressure is supplied to the oil chamber 22.
Further, from the second hydraulic pump 29, via the second hydraulic pump 29 and the second flow control valve 30 and the oil passage 3 3 in the middle of the branch point 3 3 a and the oil passage 34 connecting the second The oil pressure is supplied to the oil chamber 23 of the engine.

In the oil passage 31, a relief valve 35 is connected to a branch point 31b provided between the first hydraulic pump 27 and the branch point 31a. In the oil passage 33, a relief valve 36 is connected to a branch point 33b provided between the second hydraulic pump 29 and the branch point 33a.
The control unit that controls the operation of the CVT 5 and the engine 2 is configured by an ECU (Electronic Control Unit) 40.

  The ECU 40 includes an accelerator operation amount sensor 41 that detects an accelerator operation amount θ, a vehicle speed sensor 42 that detects a vehicle speed V that is a traveling speed of the vehicle, an engine rotation speed sensor 43 that detects a rotation speed ωe of the engine 2, and a CVT. The CVT input shaft rotational speed sensor 44 for detecting the rotational speed ωi of the input shaft 6, the CVT output shaft rotational speed sensor 45 for detecting the rotational speed ωo of the CVT output shaft 7, and the first oil chamber 22 of the hydraulic cylinder 21. A first pressure sensor 46 that detects the hydraulic pressure, a second pressure sensor 47 that detects the hydraulic pressure of the second hydraulic pressure 23 of the hydraulic cylinder 21, and signals from these sensors 41 to 47 are input to the ECU 40. It has become.

The ECU 40 outputs command signals D1, D2 to the first flow control valve 28 and the second flow control valve 30, respectively, in order to control the torque transmission force of the rollers 17, 18. Further, ECU 40, in order to control the engine output, and outputs a command signal D3 to the fuel supply amount adjusting mechanism 4 8 for adjusting the amount of fuel supplied to the engine 2. Command signal D3 is a voltage command signal for Sorenodo solenoid valve for example a throttle valve opening control of the fuel supply amount adjusting mechanism 4 8.

The ECU 40 also inputs the rotational speed ωi of the CVT input shaft 6 detected by the CVT input shaft rotational speed sensor 44 and the rotational speed ωo of the CVT output shaft 7 detected by the CVT output shaft rotational speed sensor 45. Based on the above, the gear ratio r _CVT (r _CVT = ωo / ωi) is calculated.
Further, the ECU 40 inputs the hydraulic pressure of the first oil chamber 22 detected by the first pressure sensor 46 and the hydraulic pressure of the second oil chamber 23 detected by the second pressure sensor 47, and based on these. The piston differential pressure P is calculated.

FIG. 3 is a block diagram showing a control concept. Referring to FIG. 3, a driver request conversion unit 51 and a main control unit 40 a are provided as main elements of the ECU 40, and the main control unit 40 a includes a piston differential pressure control unit 52 and an engine torque control unit 53. Including.
The driver request conversion unit 51 inputs the accelerator operation amount θ detected by the accelerator operation amount sensor 41 and the vehicle speed V detected by the vehicle speed sensor 42, and based on these inputs, a target for giving the engine 2 maximum efficiency. The engine speed ω _{e, T} and the target gear ratio r _{CVT, T} are calculated.

The piston differential pressure control unit 52 controls the piston differential pressure using feedback control of the transmission ratio. The piston differential pressure control unit 52 functions as a sensitivity reduction means for reducing the sensitivity of the transmission ratio of the CVT 5 with respect to an external rotational speed fluctuation. Specifically, the piston differential pressure control unit 52 includes a transmission ratio control unit 54 and a transmission ratio / piston differential pressure conversion unit 55.
Gear ratio control unit 54, the target speed ratio r _CVT, enter the deviation of the actual speed ratio r _CVT is _T, the feedback control of the gear ratio based on the target speed ratio r _{CVT, T} and the actual gear ratio r _CVT To do.

Speed ratio piston differential pressure converter 55 converts the transmission ratio input from the speed change ratio control unit 54 to request the piston differential pressure P _D. Specifically, the gear ratio / piston differential pressure conversion unit 55 has an actual gear ratio r _CVT , an actual piston differential pressure p, and a corrected target gear ratio r _{CVT, AT} input from the gear ratio control unit 54. Entered.
The gear ratio / piston differential pressure conversion unit 55 uses the actual gear ratio r _CVT , the actual piston differential pressure p, and the corrected target gear ratio r _{CVT, AT} , based on the following equation (1), The target gear ratio r _{CVT, AT} ) is converted into the required piston differential pressure p _D.

p _D = K _p · T _{i, est} (p, r _CVT ) / R _i (r _{CVT, AT} ) (1)
Here, K _p is a geometric constant. T _{i, est} (p, r _CVT ) is an estimated CVT input torque estimated using the actual piston differential pressure p and the actual gear ratio r _CVT . R _i (r _{CVT, AT} ) is a contact radius of the rollers 17, 18 with respect to the input disks 13, 14 obtained from the corrected target speed ratio r _{CVT, AT} .

Based on the required piston differential pressure p _D calculated by the speed ratio piston differential pressure converter 55, respectively command signal to the first flow rate control valve 28 and the second flow control valve 30 D1, D 2 is output, The piston differential pressure is controlled based on the gear ratio.
On the other hand, the engine torque control unit 53 controls the engine torque by using feedback of the engine rotation speed. Specifically, the engine torque control unit 53 includes an engine rotation speed control unit 56 and a CVT torque compensation unit 57. Engine rotational speed control section 56 receives the deviation of the actual engine rotational speed omega _e between the target engine rotational speed omega _{e, T,} based on the target engine rotational speed omega _{e, T} and the actual engine rotational speed omega _e , Feedback control of engine speed.

The CVT torque compensator 57 receives the actual gear ratio r _CVT , the actual piston differential pressure p, and the corrected engine speed ω _{e, AT} input from the engine speed controller 56.
The CVT torque compensator 57 uses the actual gear ratio r _CVT , the actual piston differential pressure p, and the corrected engine rotational speed ω _{e, AT} to calculate the required engine torque T _{e, D} based on the following equation (2). Calculate. I _e is the engine inertia.

T _{e, D} = I _e · ω _{e, AT} −T _{i, est} (p, r _CVT ) (2)
Based on the required engine torque _{Te, D} output from the CVT torque compensator 57, a command signal D3 is output to the fuel supply amount adjusting mechanism 48 to control the engine torque.
According to this preferred embodiment, converts the CVT5 torque control type comprising a variator 12 of the full toroidal type, by controlling the piston pressure difference p of CVT5 using feedback speed ratio, the torque control to the gear ratio control be able to. As a result, the sensitivity of the transmission ratio r _CVT of the _CVT 5 with respect to fluctuations in the rotational speed from the outside can be reduced. In other words, the piston differential pressure control unit 52 functions as a sensitivity lowering unit that softens the sensitivity of the transmission gear ratio with respect to fluctuations in rotational speed from the outside by control.

Since the sensitivity of the transmission ratio r _CVT of the _CVT 5 with respect to fluctuations in the rotational speed from the outside can be reduced, the transmission ratio can be stabilized and the drivability can be improved regardless of the conditions. Moreover, the controllability of torque and slip, which is an advantage of the torque control type CVT, is maintained.
In other words, in the present reference embodiment focuses on the continuous conversion of the transmission ratio control and torque control, and converts the CVT torque control type transmission ratio control. This is because the internal dynamics that convert torque control CVT to gear ratio control are more flexible and more realistically robust to control system conversion than the internal dynamics that convert gear ratio control CVT to torque control This is because it is considered.

The reference embodiment of Figures 1-3, the CVT torque control type, an example of reducing by software by controlling the sensitivity of the speed ratio to the external rotational speed variation, the embodiments of FIGS. 4 to 6 Now, an example in which the sensitivity of the gear ratio is reduced in hardware will be described.
Referring to FIG. 4, the present embodiment is different from the reference embodiment of FIG. 1 in that a first oil interposed between the first oil chamber 22 of the hydraulic cylinder 21 and the first flow control valve 28 is the first. An opening degree adjusting valve 61 and a second opening degree adjusting valve 62 interposed between the second oil chamber 23 and the second flow rate control valve 30 are provided. The first and second opening degree adjusting valves 61 and 62 reduce sensitivity so as to reduce the sensitivity of the gear ratio with respect to fluctuations in the rotational speed from the outside in the torque control type CVT 5. Function as.

Specifically, the first and second opening degree adjusting valves 61 and 62 can adjust the opening degree based on a command signal from the ECU 40. The first and second opening degree adjusting valves 61 and 62 include those that restrict the opening degree to zero.
Referring to FIG. 5, the present embodiment is different from the reference embodiment of FIG. 3 as follows. That is, instead of the driver request conversion unit 51, a driver request conversion unit 51A is provided. In the driver request conversion unit 51A, the accelerator operation amount θ detected by the accelerator operation amount sensor 41 and the vehicle speed sensor 42 are provided. Based on these, the target engine speed ω _{e, T} for giving maximum efficiency to the engine 2 is calculated, and the target CVT input torque T _{CVT, T} is calculated.

Further, a piston differential pressure control unit 71 is provided instead of the piston differential pressure control unit 52. The piston differential pressure control unit 71 includes a CVT input torque control unit 72 and a torque / piston differential pressure conversion unit 73.
The CVT input torque control unit 72 inputs the target CVT input torque T _{CVT, T} , the actual gear ratio r _CVT and the actual piston differential pressure p, and these target CVT input torque T _{CVT, T} and the actual gear ratio r _CVT. Based on the actual piston differential pressure p, a corrected target CVT input torque T _{CVT, AT} is calculated and output to the torque / piston differential pressure conversion unit 73.

In the torque / piston differential pressure conversion unit 73, the corrected target CVT input torque T _{CVT, AT} , the actual gear ratio r _CVT and the actual piston differential pressure p input from the CVT input torque control unit 72 are input, and these corrected target CVTs are input. Based on the input torque T _{CVT, AT} , the actual gear ratio r _CVT, and the actual piston differential pressure p, the required piston differential pressure p _D is calculated and output.
The torque / piston differential pressure conversion unit 73 uses the actual gear ratio r _CVT , the actual piston differential pressure p, and the corrected CVT input torque T _{CVT, AT} , and based on the following equation (3), the required engine torque P _D Is calculated.

P _D = K _p · T _{CVT, AT} / R _i (r _CVT ) (3)
Based on the required piston differential pressure p _D outputted from the piston differential pressure control unit 71, respectively command signal to the first flow rate control valve 28 and the second flow control valve 30 D1, D 2 is output, the target CVT input The piston differential pressure is controlled based on the torque.
Next, referring to FIG. 6, the ECU 40 is provided with a mode setting unit 74 and first and second signal output units 75 and 76. The mode setting unit 74 alternatively sets a torque control mode and a gear ratio control mode. Normally, the torque control mode is set, and when a predetermined condition is satisfied, the gear ratio control mode is set.

The first signal output unit 75 obtains the mode setting information from the mode setting unit 74 and outputs a command signal D4 related to the opening to the first opening adjustment valve 61. The second signal output unit 76 obtains the mode setting information from the mode setting unit 74 and outputs a command signal D5 related to the opening degree to the second opening degree adjusting valve 62.
Specifically, referring to the flowchart of FIG. 7, it is monitored in step S1 whether or not a condition for shifting to the gear ratio control mode is satisfied. When the condition for shifting to the gear ratio control mode is not satisfied (NO in step S1), the torque control mode is set (step S2), and the opening degrees of the first and second opening degree adjusting valves 61 and 62 are set. Is relatively increased (step S3).

  When the condition for shifting to the gear ratio control mode is satisfied (YES in step S1), the gear ratio control mode is set (step S4), and the first and second opening adjustment valves 61 and 62 are set. The opening is made relatively small (step S5). In step S1, the condition for shifting to the gear ratio control mode is that, for example, when the traveling state shifts from the two-wheel driving state to the four-wheel driving state, the traveling load fluctuation becomes large. That is, for example, when the travel load fluctuation amount detected by the travel load detection sensor is larger than a predetermined amount, the shift ratio control mode is entered.

In the gear ratio control mode, the opening degree of the first and second opening degree adjusting valves 61 and 62 is narrowed compared to that in the torque control mode. Thus, in the gear ratio control mode, compared to the torque control mode. The sensitivity of the transmission ratio of the CVT 5 with respect to fluctuations in rotational speed from the outside can be lowered.
In the gear ratio control mode, the opening degree of both the first and second opening degree adjusting valves 61 and 62 may be reduced as shown in FIG. 7, and the acceleration direction of the vehicle is actually known. In some cases, only one of the openings may be throttled.

According to the present embodiment, the sensitivity of the gear ratio with respect to a fluctuation in rotational speed from the outside can be reduced in hardware by narrowing the opening degree of the first and second opening degree adjusting valves 61 and 62. Therefore, complicated control is not necessary.
As shown in FIG. 8, in the gear ratio control mode, the slips of the rollers 17 and 18 are actually suppressed with either one of the first and second opening adjustment valves 61 and 62 closed. Only the end load pressure (the pressure in the third oil chamber 20) may be controlled based on the CVT input torque.

  Specifically, in step S11, it is monitored whether a condition for shifting to the gear ratio control mode is satisfied. If the condition for shifting to the gear ratio control mode is not satisfied (NO in step S11), the torque control mode is set (step S12), and the first and second opening degree adjusting valves 61 and 62 are opened. (Step S13), the pressures of the first and second oil chambers 22 and 23 of the hydraulic cylinder 21 are controlled (Step S14). That is, the piston differential pressure with respect to the rollers 17 and 18 is controlled based on the differential pressure between the first and second oil chambers 22 and 23.

  On the other hand, when the condition for shifting to the gear ratio control mode is satisfied (YES in step S11), the gear ratio control mode is set (step S15), and the first and second opening adjustment valves 61, The actual CVT is controlled so as to suppress the slipping of the rollers 17 and 18 in a state where either one of the first opening 62 is closed (step S16) and one of the first and second opening degree adjusting valves 61 and 62 is closed. Only the end load pressure (the pressure in the third oil chamber 20) is controlled based on the input torque (step S17).

According to the present embodiment, in the gear ratio control mode, slip control can be easily prevented without requiring complicated control by controlling the pressure of the end load in the third oil chamber 20.
In addition, this invention is not limited to the said embodiment, A various change can be given in the claim of this invention.

1 is a schematic diagram illustrating a schematic configuration of a vehicle to which a drive control device for a vehicle according to a reference embodiment of the present invention is applied. It is the schematic of the principal part of CVT. It is a schematic diagram of the principal part of CVT. It is a block diagram which shows schematic structure of control. 1 is a schematic diagram illustrating a schematic configuration of a vehicle to which a vehicle drive control device according to an embodiment of the present invention is applied. It is a block diagram which shows schematic structure of control of embodiment of FIG. It is a block diagram which shows the principal part of the control part of embodiment of FIG. It is a flowchart which shows the flow of control in embodiment of FIG. FIG. 5 is a flowchart showing a flow of control in the modified example of FIG. 4 as another embodiment of the present invention.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 ... Drive train, 2 ... Engine, 5 ... CVT (torque control type continuously variable transmission), 6 ... CVT input shaft, 7 ... CVT output shaft, 12 ... (Full toroidal type) variator, 13, 14 ... Input Disc (input member), 15, 16 ... Output disc (output member), 17, 18 ... Roller (intermediate transmission member), 19 ... Carriage, 20 ... Oil chamber for end load, 21 ... Hydraulic pressure for roller Cylinder, 22 ... first oil chamber, 23 ... second oil chamber, 25 ... shuttle valve, 27 ... first hydraulic pump, 28 ... first flow control valve (first control valve), 29 ... first 2 ... hydraulic pump, 30 ... second flow rate control valve (second control valve), 40 ... ECU (control unit), 52 ... piston differential pressure control unit (sensitivity reducing means), 53 ... engine torque control unit, 54 ... speed ratio control unit, 55 ... speed ratio / piston difference Conversion unit 56 ... Engine rotation speed control unit 57 ... CVT torque compensation unit 61 ... First opening adjustment valve (sensitivity reduction means) 62 ... Second opening adjustment valve (sensitivity reduction means) 71 ... Piston differential pressure control unit, 72 ... CVT input torque control unit, 73 ... Torque / piston differential pressure conversion unit, 74 ... Mode setting unit, 75 ... First signal output unit, 76 ... Second signal output unit, p ... Piston differential pressure, P _D ... Required piston differential pressure, r _CVT ... Gear ratio, r _{CVT, T} ... Target gear ratio

Claims (2)

A torque control type continuously variable motor including an input member, an output member, and an intermediate transmission member biased with a predetermined piston differential pressure between the input member and the output member, and capable of continuously changing a gear ratio. A transmission (hereinafter also referred to as CVT),
Bei example a sensitivity reduction means for reducing the sensitivity of the transmission ratio of the CVT for speed variation from the outside,
The CVT is arranged in an input disk as an input member and an output disk as an output member that are urged toward each other, and a toroidal gap formed between the two disks, and an intermediate for transmitting torque between the two disks. A roller as a transmission member, a carriage that rotatably supports the roller, and first and second oil chambers that generate a differential pressure for applying a pulling force to the disks to the roller via the carriage. A hydraulic cylinder for rollers, a first control valve for generating hydraulic pressure in the first oil chamber, a second control valve for generating hydraulic pressure in the second oil chamber, and a first oil chamber A first opening adjustment valve interposed between the first control valve and the first control valve; a second opening adjustment valve interposed between the second oil chamber and the second control valve; and an input disk. To get closer to the output disc And and a third oil chamber for end loading higher pressure of the second oil chamber hydraulic pressure is introduced,
The first and second opening degree adjusting valves function as the sensitivity lowering means by narrowing the opening degree,
The control unit controls the hydraulic pressure of the third oil chamber based on the actual CVT input torque so as to suppress the slip of the roller with one of the first and second opening degree adjusting valves closed. A vehicle drive control device having a function of Oite to claim 1, the drive control apparatus for a vehicle CVT of the torque control type which comprises a full toroidal type variator.

Images