JPH06288449A - Pulley lateral pressure control device for belt-type continuously variable transmission - Google Patents

Abstract

PURPOSE:To optimize pulley lateral pressure control by computing transmission torque on the basis of driving torque from a drive source at the engaged time of a clutch means, and computing transmission torque on the basis of the inertia torque of a drive side movable pulley at the released time of the clutch means. CONSTITUTION:When a starting clutch 5 is engaged by the signal of a controller 70 to perform power transmission, a transmission torque computing means computes transmission torque on the basis of driving torque from an engine ENG. When power transmission is not performed by the release of the clutch 5, transmission torque is computed on the basis of the inertia torque of a drive side movable pulley 11. On the basis of this computed transmission torque, a pulley lateral pressure control valve 40 controls lateral pressure control oil pressure to be supplied to the cylinders of the drive side and driven side movable pulleys 11, 16. Optimum pulley lateral pressure control is thereby performed to improve the fuel consumption of the engine. As to the transmission torque from the drive side movable pulley 11 through a V-belt, belt transmission torque is obtained only by the inertia torque corresponding to the rotational change rate of the drive side pulley 11 if the drive side cylinder internal pressure is on the low pressure side.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a belt type continuously variable transmission, in which a pulley which acts to axially press the belt from the pulley so that the torque transmitted through the belt has a desired value. The present invention relates to a device for controlling lateral pressure.

[0002]

2. Description of the Related Art Belt type continuously variable transmissions have been proposed in the past and have already been partially put into practical use for automobiles and the like. The belt type continuously variable transmission is configured by, for example, winding a metal V-belt between drive-side and driven-side movable pulleys having variable pulley widths, and controlling the drive-side and driven-side pulley widths. As a result, gear change (ratio) control is performed. In order to control the pulley width as described above, a hydraulic cylinder for applying a lateral pressure to each movable pulley half of each pulley is provided. The hydraulic cylinder controls the lateral pressure of the pulley to wind the V-belt around each pulley. The radius is controlled, and shift control is performed.

The shift control is performed by controlling the pulley side pressures on the drive side and the driven side as described above. This pulley side pressure (specifically, the low side pulley side pressure) is transmitted through the belt. It is set based on the torque that is applied. Since torque transmission through the V-belt is obtained by the frictional force between the side surface of the V-belt and the side surface of the pulley groove, the pulley side pressure is actually the torque (belt transmission torque) obtained by this frictional force. The torque is set to have a certain margin so that the torque is greater than the torque transmitted through (actual transmission torque). As a result, slippage between the belt and the pulley and a reduction in transmission efficiency are prevented.

As described above, in order to prevent the belt from slipping, it is desirable to increase the pulley side pressure by a predetermined margin. However, if the margin is excessively increased, the pulley side pressure becomes excessively high, and this pulley side pressure is increased. There is a problem that the driving power of the hydraulic pump for obtaining the hydraulic pump is increased and the fuel consumption of the engine that drives the hydraulic pump is reduced, and the durability of the belt is reduced. For this reason, it is desirable to set the pulley side pressure to a value as close as possible to a value corresponding to the actual transmission torque within a range that does not cause the belt to slip.
Conventionally, various pulley side pressure setting methods have been proposed.

For example, in Japanese Patent Publication No. 2-45062, engine output torque is calculated from the engine speed and intake negative pressure, and a reduction ratio (gear ratio) is calculated from the engine speed and driven pulley speed. However, a method for obtaining the optimum pulley side pressure from these engine output torque and reduction ratio is disclosed. Further, in Japanese Patent Laid-Open No. 63-222943, the power loss generated in the drive system is corrected from the calculated engine torque to obtain the actually transmitted engine torque, and the pulley side pressure is calculated based on the corrected engine torque. A method for obtaining is disclosed. Furthermore, the fairness 3-
Japanese Patent No. 38517 discloses a method of performing more accurate pulley side pressure control by considering inertia torque generated when the pulley rotation changes and performing inertia torque correction to calculate an accurate transmission torque even when the rotation speed changes. There is.

[0006]

In such a continuously variable transmission, a clutch for starting control (referred to as a starting clutch), a clutch for forward / reverse switching control (forward / reverse switching clutch), etc. are provided. The shift range position set by the manual operation lever etc. is N
When in the (neutral) position or the P (parking) position, these clutches are released so that power transmission from the engine is not performed. For this reason,
The conventional pulley side pressure control method is effective when the clutch is engaged, but is not suitable for the control when the clutch is disengaged.

The above-mentioned conventional pulley side pressure control method is
It can be used also when the shift range position is N or P position and the clutch is released, but when the clutch is released, the torque transmitted via the belt becomes small and the pulley side pressure may be low. However, if the conventional pulley side pressure control method is used as it is, an unnecessarily high pulley side pressure is set, and there is a problem that the fuel consumption of the engine is reduced accordingly.

The present invention has been made in view of the above problems. When the clutch is disengaged, the pulley side pressure control is configured so as to be able to perform the optimum pulley side pressure control corresponding to the transmission torque at that time. The purpose is to provide a device.

[0009]

In order to achieve such an object, in the present invention, a V-belt is wound between a drive side movable pulley connected to a drive source and a driven side movable pulley connected to an output member to drive the drive belt. The side pressure control hydraulic pressure supplied to the drive side cylinder for setting the side movable pulley width and the driven side cylinder for setting the driven side movable pulley width is controlled to perform belt side pressure control and shift control. A belt type continuously variable transmission is configured by disposing clutch means for controlling connection / disconnection of power transmission with the drive side movable pulley.

The pulley side pressure control device of the present invention supplies a transmission torque calculating means for calculating the transmission torque transmitted from the drive side movable pulley via the V-belt, and supplies it to the drive side and driven side cylinders based on this transmission torque. It is composed of a side pressure control valve that regulates the side pressure control oil pressure to be applied, and a clutch control means that controls the operation of the clutch means.
Here, the transmission torque calculation means calculates the transmission torque based on the drive torque from the drive source when the clutch means is engaged by the clutch control means to transmit power, and the clutch control means calculates the transmission torque. Is released and the power transmission is cut off, the transmission torque is calculated based on the inertia torque of the drive side movable pulley.

The transmission torque calculation means is such that, when the clutch control means disengages the clutch means and the power transmission is cut off, the control oil pressure in the drive side cylinder is lower than the control oil pressure in the driven side cylinder. The transmission torque is calculated based only on the inertia torque of the drive side movable pulley, and if the control oil pressure in the driven side cylinder is lower than the control oil pressure in the drive side cylinder, the friction torque of the inertia torque of the drive side movable pulley and the V belt. It is preferable that the transmission torque is calculated based on

[0012]

When the pulley side pressure control device having the above construction is used, the shift range position is at the D, L positions and the like, and when the clutch means is engaged, the drive transmitted through this clutch means. Pulley side pressure control is performed based on the drive torque from the source. On the other hand, when the shift range position is at the N and P positions and the clutch means is released, the torque from the drive source is not transmitted to the belt side, so that the inertia torque corresponding to the rotational change of the drive side movable pulley is transmitted. Therefore, the pulley side pressure control corresponding to this inertia torque is performed.

Since the transmission torque of the belt is determined by the internal pressure on the low pressure side of the drive-side and driven-side cylinder internal pressures, the transmission torque transmitted from the drive-side movable pulley via the V-belt is the drive-side cylinder internal pressure. When it is on the low pressure side, it can be determined based on only the inertia torque corresponding to the rotation change rate of the drive pulley, but when the driven-side cylinder internal pressure is on the low pressure side, the inertia torque of the drive side movable pulley is connected to the V belt. It is calculated based on the torque added with the friction torque.

[0014]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT A preferred embodiment of the present invention will be described below with reference to the drawings. FIG. 1 schematically shows a belt type continuously variable transmission CVT having a pulley side pressure control device of the present invention. This belt type continuously variable transmission CVT has a metal V-belt mechanism 1 arranged between an input shaft 1 and a counter shaft 2.
0, a planetary gear type forward / reverse switching mechanism 20 arranged between the input shaft 1 and the drive side movable pulley 11, a counter shaft 2 and an output side member (differential mechanism 8 or the like). It is composed of a starting clutch 5 provided. The continuously variable transmission CVT is used for vehicles, the input shaft 1 is connected to the output shaft of the engine ENG via the coupling mechanism CP, and the power transmitted to the differential gear mechanism 8 is transmitted to the left and right wheels. It

The metal V-belt mechanism 10 includes a drive side movable pulley 11 arranged on the input shaft 1 and a counter shaft 2.
It is composed of a driven side movable pulley 16 disposed above and a metal V-belt 15 wound between both pulleys 11 and 16. The drive-side movable pulley 11 is composed of a fixed pulley half body 12 rotatably arranged on the input shaft 1 and a movable pulley half body 13 that is movable relative to the fixed pulley half body 12 in the axial direction. . A drive side cylinder chamber 14 is formed on the side of the movable pulley half body 13 and surrounded by a cylinder wall 12a connected to the fixed pulley half body 12.
The hydraulic pressure supplied into the drive side cylinder chamber 14 via the oil passage 39a generates a lateral pressure that moves the movable pulley half body 13 in the axial direction.

The driven side movable pulley 16 comprises a fixed pulley half body 17 fixed to the counter shaft 2 and a movable pulley half body 18 which is axially movable relative to the fixed pulley half body 17. On the side of the movable pulley half 18,
A driven side cylinder chamber 19 is formed by being surrounded by a cylinder wall 17a connected to the fixed pulley half body 17, and the movable pulley half body is driven by the hydraulic pressure supplied into the driven side cylinder chamber 19 through an oil passage 39b. A lateral pressure is generated that moves 18 axially. Therefore, by appropriately controlling the hydraulic pressure supplied to both cylinder chambers 14 and 19 (pulley side pressure control hydraulic pressure), an appropriate pulley side pressure that does not cause slippage of the belt 15 is set and both pulleys 11 and 16 are set. The width of the pulley can be changed, and thus, the winding radius of the V-belt 15 can be changed to change the gear ratio steplessly.

The planetary gear type forward / reverse switching mechanism 20 includes a sun gear 21 connected to the input shaft 1, a carrier 22 connected to the fixed pulley half 12, and a ring gear 23 that can be fixedly held by a reverse brake 27. It comprises a forward clutch 25 capable of connecting the sun gear 21 and the ring gear 23. When the forward clutch 25 is engaged, all gears 21, 2
2, 23 rotate integrally with the input shaft 1, and the drive pulley 11 is driven in the same direction (forward direction) as the input shaft 1. When the reverse brake 27 is engaged, the ring gear 23 is fixedly held, so that the carrier 22 is driven in the direction opposite to the sun gear 21, and the drive pulley 11 is opposite to the input shaft 1 (reverse direction). Driven to.

When both the forward clutch 25 and the reverse brake 27 are released, the power transmission via the forward / reverse switching mechanism 20 is cut off, and the power between the engine ENG and the drive side drive pulley 11 is cut off. Communication will not occur. As can be seen from this, in this embodiment, the forward clutch 25 and the reverse brake 27 correspond to the clutch means in the claims.

The operation of the forward clutch 25 and the reverse brake 27 is controlled by a manual valve (not shown) which is operated according to the operation of a manual shift lever (not shown). For example, the forward clutch 25 is engaged when the manual shift lever is in the forward drive position such as the D and L positions, and the reverse brake 27 is engaged when the manual shift lever is in the reverse drive position such as the R position. When the manual shift lever is in the neutral position such as N and P, the forward clutch 25
Also, the reverse brake 27 is released.

The starting clutch 5 is a clutch for controlling the on / off of power transmission between the counter shaft 2 and the output side member, and when this is engaged (on), power transmission between the two is possible. . Therefore, when the start clutch 5 is on, the engine output changed by the metal V-belt mechanism 10 is transmitted to the differential mechanism 8 through the gears 6a, 6b, 7a, 7b, and the differential mechanism 8 causes the left and right wheels ( (Not shown) and transmitted. When the starting clutch 5 is off, this power transmission cannot be performed and the transmission is in the neutral state.

The starting clutch 5 is engagement-controlled when the manual shift lever is in the traveling side position such as D, L, and R when the vehicle is started, and the manual shift lever is set to N, P.
When it is in the neutral position, it is released. As can be seen from this configuration, the starting clutch 5 of this embodiment is the transmission CV.
Although it is arranged on the output side of T, the starting clutch 5 may be arranged on the input side, and at this time, it corresponds to the clutch means in the claims.

The operation control of the starting clutch 5 is performed by a clutch control valve 75 which is operated in response to a signal from the controller 70. The clutch control valve 75 controls the operating hydraulic pressure supplied to the starting clutch 5 via the oil passages 31a and 31b, and controls the operation of the starting clutch 5. Although not shown, a forward / reverse control valve that controls the operation of the forward clutch 25 and the reverse brake 27 is also provided, and the controller 70 also controls the operation of the forward / reverse control valve.

In order to perform such various controls, the controller 70 is supplied with an engine speed Ne signal and an engine intake negative pressure PB signal from an electronic control unit ECU which controls the operation of the engine ENG. Further, detection from a shift range position detector 66 that detects a shift range position based on a detection signal from an air conditioner operation detector 65 that detects the operation of the air conditioner AC and a manual shift lever position ATP (or a manual shift valve position). A signal is sent.

On the other hand, the control of the hydraulic pressure supplied to the drive side and driven side cylinder chambers 14 and 19 (pulley side pressure control hydraulic pressure) is controlled in accordance with a control signal from the controller 70. The pulley side pressure control valve 40 and the shift control valve are operated. 50. These pulley side pressure control valves 40
The specific structure of the shift control valve 50 is shown in FIG. In this figure, the high pressure regulator valve 41
The pulley side pressure control valve 40 is constituted by the low pressure regulator valve 43 and the high / low pressure control valve 45, and the shift control valve 51 and the shift valve 5
The shift control valve 50 is constituted by 3 and.

Hereinafter, description will be made with reference to FIG. In addition, in FIG. 2, a cross indicates that the portion is connected to the drain. The oil discharged from the hydraulic pump 30 is supplied to the high pressure control valve 41 via the oil passage 32 and is supplied to the reducing valve 58 via the oil passage 36. In the reducing valve 58, a line pressure PMOD having a substantially constant hydraulic pressure is produced, and the hydraulic oil having this line pressure is supplied to the high / low pressure control valve 45 and the shift control valve 51 via the oil passages 37a and 37b. Supply.

The high / low pressure control valve 45 has a linear solenoid 45a, and the linear solenoid 4a is controlled by controlling the current supplied to the linear solenoid 45a.
The pressing force acting on the spool 45b from 5a is controlled, the line pressure PMOD supplied from the oil passage 37a is regulated, and the control back pressure PHLC corresponding to this pressing force is supplied to the oil passage 35. The control back pressure PHLC is supplied to the right end oil chambers 41b and 43b of the high pressure and low pressure control valves 41 and 43, and the control back pressure PHLC acts to press the spools 41a and 43a to the left.

The high-pressure control valve 41 receives the control back pressure PHLC and regulates the operating hydraulic pressure supplied from the pump 30 through the oil passage 32 in accordance with the control back pressure PHLC to control the high side pressure control pressure PH. Supply to the oil passages 33a and 33b. This high side pressure control pressure PH is applied to the oil passage 33.
It is supplied to the shift valve 53 via a and to the low pressure control valve 43 via the oil passage 33b. The low pressure control valve 43 has a control back pressure PHLC.
In response to this, the high side pressure control pressure PH supplied from the oil passage 33b is adjusted, and the low side pressure control pressure PL is supplied to the oil passage 3b.
Supply to 4. The low side pressure control pressure PL is supplied to the shift valve 53 via oil passages 34a and 34b branched from the oil passage 34.

On the other hand, the shift control valve 51 is
The linear solenoid 51a has a linear solenoid 51a.
By controlling the energizing current to a, the pressing force applied from the linear solenoid 51a to the spool 51b is controlled, and the line pressure PMOD supplied from the oil passage 37b is adjusted to adjust the shift control pressure corresponding to this pressing force. Supply PSV to the oil passage 38. This shift control pressure PSV is supplied to the left end oil chamber 53b of the shift valve 53 and acts so as to press the spool 53a to the right.

Here, the spool 53a is pressed to the left by the spring 53c from the right end side, and is moved to a position where the shift control pressure PSV and the pressing force of the spring 53c are balanced. Therefore, by controlling the shift control pressure PSV by the shift control valve 51, the position control of the spool 53a of the shift valve 53 can be performed. Thereby, the high and low side pressure control pressures PH and PL are appropriately distributed and supplied to the drive side and driven side cylinder chambers 14 and 19,
Shift control can be performed.

When the shift control is performed in this manner, the low side pressure control pressure PL regulated by the low pressure regulator valve 43 is an optimum pulley for transmitting a predetermined torque without causing a belt slip. It is set to be the lateral pressure. The setting control of the pulley side pressure is performed by the controller 70. This control will be described with reference to the block diagram of FIG. 3 and the flowcharts of FIGS. 4 and 5.

In this control, as shown in FIGS. 3 and 4, first, the engine speed Ne and the engine intake negative pressure P
B, drive pulley rotation speed NDR and driven pulley rotation speed NDN are read (step S1, block B
1, B2, B11, B12). As shown in FIG. 1, the engine ENG includes an electronic control unit ECU that controls the operation of the engine ENG.
Is provided, and the engine speed N is set for this control.
Since e and the engine intake negative pressure PB are detected and used, the detected value is directly input to the controller 70. Further, the low side pressure control pressure PLCMD set in the previous flow (this flow is repeated every predetermined control cycle) is read (step S2, block B15).

Next, the control from steps S2 to S7 is performed to determine the engine output torque TE according to the presence / absence of engine fuel cut. Specifically, when the engine fuel cut is not performed, in step S1 from the engine drive torque TEPB map which is measured in correspondence with the engine speed Ne and the intake negative pressure PB and is shown in FIG. An engine drive torque TEPB corresponding to the read current engine speed Ne and intake negative pressure PB is obtained and set as an engine output torque TE (steps S4, S5, blocks B3, B5).

On the other hand, when the engine fuel cut is being performed, the engine friction torque FE measured and set as shown in FIG. 7 corresponding to the engine speed Ne.
From the map, the engine friction torque FE corresponding to the current engine speed Ne (this torque is a torque that works in the opposite direction to the normal engine drive torque) is obtained, and this is set as the engine output torque TE (step S6, S7, blocks B4, B5).

Next, from the drive side and driven side pulley rotation speeds NDR, NDN, the reduction ratio i (=
NDR / NDN) is calculated (step S8, block B16). Further, the rate of change DNDR of the drive side pulley rotation speed NDR is obtained (step S9, block B1).
3). This value is obtained, for example, from the difference between the drive-side pulley rotation speed in the previous flow and the current drive-side pulley rotation speed. Rotational change rate DND obtained in this way
From R, the inertia torque IE of the engine-side inertia system and the inertia torque IDR of the drive-side pulley inertia system necessary for causing such a rotation change are calculated (steps S10, S11, block B14). In addition,
As shown in FIG. 9, this calculation is performed using an inertia torque map which is calculated and stored in advance according to the rotational change rate of each inertia.

Further, the friction torque FPU necessary for driving the oil pump corresponding to the drive side pulley rotation speed NDR and the low side pressure control pressure PLCMD in the previous flow.
MP is obtained (step S12, block B17). This pump drive friction torque FPUMP is also obtained using a map previously measured and stored corresponding to the low side pressure control pressure PL, as shown in FIG.

Then, the belt drive friction torque FBLT generated in this state is calculated using the reduction ratio i and the previous low side pressure control pressure PLCMD (step S).
13, block B18). This belt drive friction torque FBLT is also obtained using a map previously measured and stored corresponding to the low side pressure control pressure PL and the reduction ratio i, as shown in FIG.

Further, the air conditioner drive friction torque FAC is obtained based on the signal from the air conditioner operation detector 65 (step S14, block B8). The air conditioner drive friction torque FAC differs depending on whether the air conditioner is on or off. As shown in FIG. 8, when the air conditioner is off, the friction torque FAC is substantially zero, and when it is on, a predetermined friction torque is obtained.

The belt transmission torque TBLT is calculated using the various torques obtained as described above (step S15, blocks B20 to B23). Belt transmission torque T
The BLT calculation flow is shown in detail in Fig. 5. In this calculation, first, the shift range position detected by the shift range position detector 66 in step S21 is N
Alternatively, it is determined whether or not it is in the P position. The shift range position is a position other than N and P positions (neutral position) (D, L,
If it is in the traveling side position such as R), the process proceeds to step S22, and it is determined whether the hydraulic pressure of the driven side cylinder chamber is on the low pressure side, that is, whether the driven side cylinder chamber pressure is lower than the drive side cylinder chamber pressure. To be judged.

When the internal pressure of the driven side cylinder chamber is on the high pressure side, the internal pressure of the drive side cylinder chamber becomes the low side pressure control pressure PL, so that it is not necessary to consider the influence of the belt drive friction torque FBLT.
23, the actual transmission torque TIN transmitted from the drive pulley via the belt is calculated by the formula TIN = TE-IE-IDR-FPUMP-FAC. On the other hand, when the internal pressure of the driven side cylinder chamber is on the low pressure side, it is necessary to consider the belt drive friction torque FBLT, so the routine proceeds to step S24, and the formula TIN = TE-IE-IDR-FPUMP-FBLT-FAC The actual transmission torque TIN transmitted from the pulley to the belt is calculated.

When the shift range position is at the N or P position, the routine proceeds to step S25, where it is determined whether or not the hydraulic pressure in the driven side cylinder chamber is on the low pressure side, that is, the driven side cylinder chamber pressure is lower than the drive side cylinder chamber pressure. It is determined whether or not. At the N and P positions, since the power transmission is cut off by the forward / reverse switching mechanism 20 and the starting clutch 5, the actual transmission torque TIN is determined by the inertia torque IDR of the drive-side pulley inertia system. However, there is a difference in whether or not the belt drive friction torque FBLT needs to be taken into consideration depending on whether the internal pressure of the driven side cylinder chamber is the high pressure side or the low pressure side. Therefore, this determination is made in step S25.

When the internal pressure of the driven side cylinder chamber is on the high pressure side, the belt drive friction torque FBL.
Since it is not necessary to consider T, the process proceeds to step S26, and the actual transmission torque TIN transmitted from the drive pulley via the belt is calculated by the formula TIN = -IDR. On the other hand, when the internal pressure of the driven side cylinder chamber is on the low pressure side, it is necessary to consider the belt drive friction torque FBLT, so the routine proceeds to step S27, and the formula TIN = -IDR-FBLT is used to transmit from the drive side pulley through the belt. The actual transmission torque TIN is calculated.

As described above, steps S21 to S27 (block B)
When the actual transmission torque TIN is calculated in 20), it is then determined in step S28 whether the actual transmission torque TIN is positive or negative. When the actual transmission torque TIN is positive, the first safety factor SFP (> 1.
0) is set as the safety factor SF, and when the actual transmission torque TIN is negative, the second safety factor SFM (> SFP) larger than the first safety factor SFP is set as the safety factor SF (block B21). Then, the absolute value of the actual transmission torque TIN is multiplied by this safety factor SF to obtain the belt transmission torque TIN.
BLT is calculated (step S31, block B2)
2).

The belt transmission torque TBLT calculated in this way is a torque having a predetermined margin, that is, a safety factor, with respect to the actual transmission torque TIN that actually acts on the drive side pulley. The safety factor is set to a large value when the actual transmission torque TIN is negative. Therefore, if the pulley side pressure control is performed so that the belt transmission torque TBLT calculated in this way is obtained, optimum control can be performed.

When the belt transmission torque TBLT is calculated as described above, step S16 (block B)
In step 23), the target low side pressure control pressure PLCMD is calculated corresponding to the speed reduction ratio i at that time. It should be noted that this calculation is obtained from a preset map corresponding to the belt transmission torque TBLT and the reduction ratio i, as shown in FIG.

The target low side pressure control pressure PLCMD thus obtained may be generated by the low pressure regulator valve 43. However, as can be seen from the circuit configuration of FIG. The current control determines the hydraulic pressure controlled by the high pressure regulator valve 41 and the low pressure regulator valve 43. Therefore, as shown in FIGS. 13, 14 and 15, the relationship between the energization current of the linear solenoid 45a and the low side pressure control pressure PL, the energization current of the linear solenoid 45a and the high / low pressure control valve 45.
Relationship between the control back pressure PHLC created by the control back pressure PHLC and the control back pressure PHLC and the high side pressure and the low side pressure control pressure P
The relationship between H and PL is preset and stored. And
The energizing current of the linear solenoid 45a is determined based on these maps stored and stored, and this energizing current control is performed.

When the high side pressure and the low side pressure control pressures PH and PL are set in this way, the shift valve 5
As shown in FIG. 16, the control for appropriately distributing these to the drive and driven-side cylinder chambers 14 and 19 is performed by 3, and the shift control is performed.

[0047]

As described above, according to the present invention,
The transmission torque calculating means for calculating the transmission torque transmitted from the drive side movable pulley via the V-belt is the drive torque from the drive source when the clutch control means is engaged with the clutch means to transmit power. The transmission torque is calculated based on, and when the clutch control means releases the clutch means and the power transmission is cut off,
The transmission torque is calculated based on the inertia torque of the drive side movable pulley. Therefore, for example, if the manual shift lever is at the traveling side position (D, L, R
Position etc.) and the clutch means is engaged, and when the manual shift lever is positioned at the neutral position (N, P position etc.) and the clutch means is released, the actual transmission torque at that time is It is possible to improve the fuel efficiency of the engine by performing the optimum matched pulley side pressure control.

Since the transmission torque of the belt is determined by the internal pressure on the low pressure side of the drive-side and driven-side cylinder internal pressures, the transmission torque transmitted from the drive-side movable pulley via the V-belt is the drive-side cylinder internal pressure. When it is on the low pressure side, it is obtained based on only the inertia torque corresponding to the rotation change rate of the drive side pulley, and when the driven side cylinder internal pressure is on the low pressure side,
It is calculated based on the torque obtained by adding the friction torque of the V belt to the inertia torque of the drive side movable pulley.

[Brief description of drawings]

FIG. 1 is a schematic diagram showing a configuration of a belt type continuously variable transmission having a pulley side pressure control device of the present invention.

FIG. 2 is a hydraulic circuit diagram showing the configuration of this pulley side pressure control device.

FIG. 3 is a block diagram showing control contents by the pulley side pressure control device.

FIG. 4 is a flowchart showing the contents of control by this pulley side pressure control device.

FIG. 5 is a flowchart showing the contents of control by this pulley side pressure control device.

FIG. 6 is a graph showing an engine drive torque map.

FIG. 7 is a graph showing an engine friction torque map.

FIG. 8 is a graph showing an air conditioner drive friction torque map.

FIG. 9 is a graph showing an inertia torque map.

FIG. 10 is a graph showing a pump drive friction torque map.

FIG. 11 is a graph showing a belt drive friction torque map.

FIG. 12 is a graph showing a target low side pressure control pressure map.

FIG. 13 is a graph showing the relationship between the low side pressure control pressure and the energization current of the linear solenoid.

FIG. 14 is a graph showing a relationship between a current supplied to a linear solenoid and a control back pressure.

FIG. 15 is a graph showing the relationship between the control back pressure and the high and low side pressure control pressures.

FIG. 16 is a graph showing a drive and driven-side cylinder chamber pressure corresponding to shift control.

[Explanation of symbols]

 1 Input Shaft 2 Counter Shaft 5 Starting Clutch 8 Differential Mechanism 10 Metal V Belt Mechanism 11 Drive Side Movable Pulley 15 Metal V Belt 16 Driven Side Movable Pulley 20 Forward / Reverse Switching Mechanism 30 Hydraulic Pump 40 Pulley Side Pressure Control Valve 41 High Pressure Regulator Valve 43 Low Pressure Regulator valve 45 High and low pressure control valve 50 Shift control valve 51 Shift control valve 53 Shift valve 58 Reducing valve 65 Air conditioner operation detector 66 Shift range position detector 70 Controller

Claims (2)

[Claims] 1. A drive-side movable pulley connected to a drive source, a driven-side movable pulley connected to an output member, a V-belt wound between these drive-side and driven-side movable pulleys, and a pulley of the drive-side movable pulley. A drive side cylinder for setting the width, a driven side cylinder for setting the pulley width of the driven side movable pulley, and a clutch means for controlling connection / disconnection of power transmission between the drive source and the drive side movable pulley. In the belt type continuously variable transmission described above, transmission torque calculating means for calculating the transmission torque transmitted from the drive side movable pulley via the V belt,
The transmission torque calculation includes: a side pressure control valve that regulates a side pressure control oil pressure supplied to the drive side cylinder and the driven side cylinder based on the transmission torque; The means calculates a transmission torque based on a driving torque from the drive source when the clutch control means is engaged with the clutch means to transmit the power, and the clutch control means causes the clutch means to operate. Is released and the power transmission is cut off, the transmission torque is calculated based on the inertia torque of the drive side movable pulley. Pulley side pressure control of the belt type continuously variable transmission characterized by the above-mentioned. apparatus. 2. The transmission torque calculation means controls the control hydraulic pressure in the drive side cylinder within the driven side cylinder when the clutch control means releases the clutch means to interrupt the power transmission. If the pressure is lower than the hydraulic pressure, the transmission torque is calculated based only on the inertia torque of the drive side movable pulley. If the control hydraulic pressure in the driven side cylinder is lower than the control hydraulic pressure in the drive side cylinder, the drive side movable 2. The pulley side pressure control device for a belt type continuously variable transmission according to claim 1, wherein the transmission torque is calculated based on the inertia torque of the pulley and the friction torque of the V belt.