JP3102723B2 - Pulley side pressure control device for belt type continuously variable transmission - Google Patents

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, and more particularly, to a pulley which acts to press a belt from a pulley in an axial direction so that a torque transmitted through the belt becomes 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, and some of them have already been put into practical use for automobiles and the like. The belt-type continuously variable transmission is configured, for example, by winding a metal V-belt between a drive side and a driven side movable pulley, each of which has a variable pulley width, and controls the drive side and driven side pulley widths. As a result, a shift (ratio) control is performed. In order to control the pulley width, a hydraulic cylinder that applies a side pressure to the movable pulley half of each pulley is provided, and the hydraulic cylinder controls the pulley side pressure to wind the V belt around each pulley. The radius is controlled, and shift control is performed.

Transmission control is performed by controlling the drive-side and driven-side pulley-side pressures as described above. The pulley-side pressure (specifically, the low-pressure-side pulley-side pressure) is transmitted via a belt. It is set based on the torque to be applied. Since the torque transmission via 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 torque (belt transmission torque) obtained by the frictional force is actually the pulley side pressure. Is set with a certain margin so as to be larger than the torque transmitted through the relay (actual transmission torque). This prevents the transmission efficiency from being reduced due to slippage between the belt and the pulley.

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

For example, Japanese Patent Publication No. 2-45062 discloses that an engine output torque is calculated from an engine speed and an intake negative pressure, and a reduction ratio (speed ratio) is calculated from the engine speed and a driven pulley speed. In addition, a method of obtaining an optimum pulley side pressure from the engine output torque and the reduction ratio is disclosed. Japanese Patent Application Laid-Open No. 63-222943 discloses that an engine torque actually transmitted is obtained by correcting a power loss generated in a drive system from a calculated engine torque, and a pulley side pressure is determined based on the corrected engine torque. Is disclosed. In addition,
Japanese Patent No. 38517 discloses a method of performing an inertia torque correction in consideration of an inertia torque generated at the time of a pulley rotation change, calculating an accurate transmission torque at the time of a rotation change, and performing more accurate pulley side pressure control. I have.

[0006]

By the way, in such a continuously variable transmission, a clutch for starting control (referred to as a starting clutch), a clutch for controlling forward / reverse switching (forward / backward switching clutch) and the like 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 disengaged 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 control when the clutch is disengaged.

The conventional pulley side pressure control method described above is
It can be used when the clutch is disengaged at the shift range position N or P, but when the clutch is disengaged, the torque transmitted via the belt is small and the pulley side pressure may be low. However, if the conventional pulley side pressure control method is used, an unnecessarily high pulley side pressure will be set, and there is a problem that the fuel efficiency of the engine is reduced accordingly.

The present invention has been made in view of the above problems, and has a structure in which when the clutch is disengaged, the pulley side pressure control is configured to perform optimum pulley side pressure control corresponding to the transmission torque at that time. It is intended to provide a device.

[0009]

According to the present invention, a V-belt is wound around a drive-side movable pulley connected to a drive source and a driven-side movable pulley connected to an output member. The belt side pressure control and the speed change control are performed by controlling the side pressure control oil 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. A belt-type continuously variable transmission is configured by disposing clutch means for controlling connection and disconnection of the power transmission between the drive side movable pulley and the drive side movable pulley.

A pulley-side pressure control device according to the present invention calculates a transmission torque transmitted from a drive-side movable pulley via a V-belt, and supplies the torque to a drive-side and driven-side cylinder based on the transmission torque. And a clutch control means for controlling the operation of the clutch means.
Here, the transmission torque calculating means includes a drive torque from the drive source, an inertia torque of an output rotary member of the drive source, and a drive side movable pulley when the clutch control means engages the clutch means to transmit power. The transmission torque is calculated based on the inertia torque of the drive pulley, and when the clutch control means releases the clutch means and the power transmission is cut off, the drive torque from the drive source is not used and the transmission torque is calculated based on the inertia torque of the drive side movable pulley. To calculate the transmission torque.

When the power transmission is cut off by the clutch control means when the clutch means is disengaged and the power transmission in the drive side cylinder is lower than the control hydraulic pressure in the driven side cylinder, the transmission torque calculation means is provided. The transmission torque is calculated based only on the inertia torque of the drive side movable pulley. If the control oil pressure in the driven side cylinder is lower than the control oil pressure in the drive side cylinder, the inertia torque of the drive side movable pulley and the friction torque of the V-belt It is preferable to calculate the transmission torque based on the following equation.

[0012]

When the pulley side pressure control device having the above structure is used, when the shift range position is at the D or L position and the clutch means is engaged, the drive transmitted through the clutch means is provided. Pulley-side pressure control is performed based on the driving torque from the source, the inertia torque of the output rotating member of the driving source, and the inertia torque of the drive-side movable pulley. On the other hand, the shift range positions are at the N and P positions,
When the clutch means is released, the torque from the drive source and the inertia torque of the output rotation member of the drive source are not transmitted to the belt side, so that only the inertia torque corresponding to the rotation change of the drive-side movable pulley is transmitted. Therefore, the pulley side pressure control corresponding to the inertia torque is performed without using the driving torque from the driving source.

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

[0014]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Preferred embodiments 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. The belt-type continuously variable transmission CVT includes a metal V-belt mechanism 1 disposed between an input shaft 1 and a counter shaft 2.
0, a planetary gear type forward / reverse switching mechanism 20 disposed between the input shaft 1 and the drive-side movable pulley 11, and a transmission mechanism disposed between the counter shaft 2 and an output member (such as the differential mechanism 8). The start clutch 5 is provided. The continuously variable transmission CVT is used for a vehicle, the input shaft 1 is connected to an output shaft of an engine ENG via a coupling mechanism CP, and the power transmitted to the differential mechanism 8 is transmitted to left and right wheels. You.

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

The driven-side movable pulley 16 comprises a fixed pulley half 17 fixed to the counter shaft 2 and a movable pulley half 18 movable relative to the fixed pulley half 17 in the axial direction. On the side of the movable pulley half 18,
A driven cylinder chamber 19 is formed surrounded by a cylinder wall 17a connected to the fixed pulley half 17, and the movable pulley half is driven by hydraulic pressure supplied into the driven cylinder chamber 19 via an oil passage 39b. A lateral pressure is generated which moves the shaft 18 in the axial direction. For this reason, by appropriately controlling the hydraulic pressure (pulley-side pressure control hydraulic pressure) supplied to the two cylinder chambers 14 and 19, an appropriate pulley-side pressure that does not cause slippage of the belt 15 is set and the two pulleys 11 and 16 are set. Of the pulley can be changed, whereby the winding radius of the V-belt 15 can be changed to change the speed ratio in a stepless manner.

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, a ring gear 23 which can be fixed and held by a reverse brake 27, and 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 the gears 21 and
The drive pulley 11 is driven in the same direction as the input shaft 1 (forward direction). When the reverse brake 27 is engaged, the ring gear 23 is fixed and held, so that the carrier 22 is driven in a direction opposite to that of the sun gear 21, and the drive pulley 11 is driven in a direction opposite to the input shaft 1 (reverse direction). Is driven.

When the forward clutch 25 and the reverse brake 27 are both 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. No transmission occurs. 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 in response to the operation of a manual shift lever (not shown). For example, when the manual shift lever is at a forward position such as the D or L position, the forward clutch 25 is engaged, and when the manual shift lever is at a reverse position such as the R position, the reverse brake 27 is engaged. When the manual shift lever is at a neutral position such as N or P, the forward clutch 25
And the reverse brake 27 are both released.

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

The start clutch 5 is engaged and controlled when the manual shift lever is started at a traveling position such as D, L, R or the like.
Etc. are released when in the neutral position. As can be seen from this configuration, the starting clutch 5 of the present embodiment has the transmission CV
Although it is arranged on the output side of T, the starting clutch 5 may be arranged on the input side, in which case it corresponds to the clutch means in the claims.

The operation of the starting clutch 5 is controlled 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 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 / backward control valve for controlling 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 / backward control valve.

The controller 70 receives an engine speed Ne signal and an engine intake negative pressure PB signal from an electronic control unit ECU for controlling the operation of the engine ENG so that various controls can be performed. Further, a shift range position detector 66 for detecting a shift range position based on a detection signal from an air conditioner operation detector 65 for detecting 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 by a pulley-side pressure control valve 40 and a shift control valve which are operated in response to a control signal from a controller 70. 50. These pulley side pressure control valves 40
FIG. 2 shows a specific configuration of the transmission control valve 50. In this figure, the high pressure regulator valve 41
, The low pressure regulator valve 43 and the high / low pressure control valve 45 constitute a pulley side pressure control valve 40, and the shift control valve 51 and the shift valve 5
3 constitute a transmission control valve 50.

Hereinafter, description will be made with reference to FIG. In FIG. 2, the mark “x” means that the part 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 generated, and the hydraulic oil having the 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 by controlling the current flowing to the linear solenoid 45a,
By controlling the pressing force acting on the spool 45b from 5a, the line pressure PMOD supplied from the oil passage 37a is adjusted, 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 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 adjusts the working oil pressure supplied from the pump 30 via the oil passage 32 in accordance with the control back pressure PHLC, and adjusts the high side pressure control pressure PH. The oil is supplied to the oil passages 33a and 33b. This high side pressure control pressure PH is
a, and is supplied to the low-pressure control valve 43 via an oil passage 33b. The low pressure control valve 43 is a control back pressure PHLC
In response, the high side pressure control pressure PH supplied from the oil passage 33b is adjusted, and the low side pressure control pressure PL is adjusted to the oil passage 3
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
A linear solenoid 51a having a linear solenoid 51a;
a of the linear solenoid 51a to control the pressing force acting on the spool 51b. The line pressure PMOD supplied from the oil passage 37b is adjusted to control the shift control pressure corresponding to the pressing force. The PSV is supplied 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 to press the spool 53a rightward.

Here, the spool 53a is pressed leftward from the right end by a spring 53c, and is moved to a position where the shift control pressure PSV and the pressing force of the spring 53c are balanced. Therefore, if the shift control pressure PSV is controlled by the shift control valve 51, the position of the spool 53a of the shift valve 53 can be controlled. As a result, 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, respectively.
Shift control can be performed.

When shifting control is performed in this manner, the low-side pressure control pressure PL, which is set by the low-pressure regulator valve 43, is adjusted to an optimum pulley for transmitting a predetermined torque without causing belt slippage. It is set to be the side pressure. The control for setting 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.

As shown in FIGS. 3 and 4, this control is performed by first controlling the engine speed Ne and the engine intake negative pressure P.
B, the drive pulley rotation speed NDR and the driven pulley rotation speed NDN are read (step S1, block B
1, B2, B11, B12). As shown in FIG. 1, an electronic control unit ECU for controlling the operation of the engine ENG is provided.
Is provided, and the engine speed N
Since e and the engine intake negative pressure PB are detected and used, the detected values are taken into the controller 70 as they are. 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 of steps S2 to S7 is performed to determine the engine output torque TE according to the presence or absence of the engine fuel cut. Specifically, when the engine fuel cut is not performed, the engine drive torque TEBP map measured as shown in FIG. 6 corresponding to the engine speed Ne and the intake negative pressure PB is used in step S1. An engine drive torque TEPB corresponding to the read current engine speed Ne and the intake negative pressure PB is obtained and set as the 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 according to the engine speed Ne and set as shown in FIG.
From the map, an engine friction torque FE (this torque is a torque acting in a direction opposite to the normal engine driving torque) corresponding to the current engine speed Ne is obtained, and set as an engine output torque TE (step S1). S6, S7, blocks B4, B5).

Next, the reduction ratio i (=) between the pulley rotation speeds NDR and NDN on the drive side and the driven side is used.
NDR / NDN) is calculated (step S8, block B16). Further, a change rate 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. The rotation change rate DND thus obtained
From R, an inertia torque IE of the engine-side inertial system and an inertia torque IDR of the drive-side pulley inertia system required to cause such a rotation change are calculated (steps S10, S11, block B14). In addition,
This calculation is performed using an inertia torque map that is calculated and stored in advance according to the rotational change rate of each inertia, as shown in FIG.

The friction torque FPU required 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). As shown in FIG. 10, the pump drive friction torque FPUMP is also obtained using a map previously measured and stored in correspondence with the low side pressure control pressure PL.

Then, using the reduction ratio i and the previous low side pressure control pressure PLCMD, a belt driving friction torque FBLT generated in this state is calculated (step S).
13, block B18). As shown in FIG. 11, the belt drive friction torque FBLT is also obtained by using a map which is measured and stored in advance corresponding to the low side pressure control pressure PL and the reduction ratio i.

Further, based on a signal from the air conditioner operation detector 65, an air conditioner driving friction torque FAC is obtained (step S14, block B8). The air conditioner driving 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 almost zero, and when the air conditioner 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 calculation flow of BLT is shown in detail in FIG. In this calculation, first, in step S21, the shift range position detected by the shift range position
Alternatively, it is determined whether or not it is at the P position. Shift range position other than N, P position (neutral position) (D, L,
R), the process proceeds to step S22, where it is determined whether the hydraulic pressure of the driven cylinder chamber is on the low pressure side, that is, whether the driven cylinder pressure is lower than the drive cylinder pressure. Is determined.

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, and it is not necessary to consider the influence of the belt drive friction torque FBLT.
Proceeding to 23, the actual transmission torque TIN transmitted from the drive pulley via the belt is calculated by the equation 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. Therefore, the process proceeds to step S24, and the equation TIN = TE−IE−IDR−FPUMP−FBLT−FAC is used. The actual transmission torque TIN transmitted from the pulley via the belt is calculated.

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

When the internal pressure of the driven cylinder chamber is on the high pressure side, the belt driving 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 equation 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. Therefore, the process proceeds to step S27, and is transmitted from the drive side pulley via the belt by the equation TIN = −IDR−FBLT. The calculated actual transmission torque TIN is calculated.

Steps S21 to S27 (block B
In step 20), when the actual transmission torque TIN is calculated, it is next 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, a 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 the 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 manner is a torque having a predetermined margin, that is, a safety factor, with respect to the actual transmission torque TIN actually acting on the drive 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 as to obtain the belt transmission torque TBLT calculated as described above, the optimum control can be performed.

When the belt transmission torque TBLT is calculated as described above, step S16 (block B
Proceeding to 23), the target low-side pressure control pressure PLCMD is obtained in accordance with the speed reduction ratio i at that time. This calculation is obtained from a map set in advance corresponding to the belt transmission torque TBLT and the reduction ratio i, as shown in FIG.

The target low side pressure control pressure PLCMD obtained in this manner may be generated by the low pressure regulator valve 43. As can be seen from the circuit configuration of FIG. 2, the energization of the linear solenoid 45a of the high / low pressure control valve 45 is performed. The current control determines the hydraulic pressure controlled by the high-pressure regulator valve 41 and the low-pressure regulator valve 43. For this reason, as shown in FIGS. 13, 14 and 15, the relationship between the energizing current of the linear solenoid 45a and the low side pressure control pressure PL, the energizing current of the linear solenoid 45a and the high / low pressure control valve 45
Between the control back pressure PHLC and the high side pressure and low side pressure control pressure P
The relationship between H and PL is set and stored in advance. And
Based on the map stored and set, the current supplied to the linear solenoid 45a is determined, and the current supplied is controlled.

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, and the shift control is performed.

[0047]

As described above, according to the present invention,
The transmission torque calculation means for calculating the transmission torque transmitted from the drive-side movable pulley via the V-belt includes a drive torque from a drive source when the clutch means is engaged and the power transmission is performed. The transmission torque is calculated based on the inertia torque of the output rotation member of the drive source and the inertia torque of the drive side movable pulley, and when the clutch means is released by the clutch control means and power transmission is cut off, the drive side movable pulley is operated. The transmission torque is calculated on the basis of the inertia torque. Therefore, for example, when the manual shift lever is located at the traveling side position (D, L, R position, etc.) and the clutch means is engaged, and when the manual shift lever is located at the neutral position (N, P position, etc.). When the clutch means is released, the optimal pulley side pressure control that matches the actual transmission torque at that time is performed, and the fuel efficiency of the engine can be improved.

Since the transmission torque of the belt is determined by the internal pressure of the low pressure side of the internal pressures of the drive side and driven side cylinders, the transmission torque transmitted from the drive side movable pulley via the V-belt is the internal pressure of the drive side cylinder. When it is on the low pressure side, it is obtained based only on 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 determined 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 the 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 a configuration of the pulley side pressure control device.

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

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

FIG. 5 is a flowchart showing control contents by the 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 driving friction torque map.

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

FIG. 10 is a graph showing a pump driving 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 a relationship between a low side pressure control pressure and a current supplied to a linear solenoid.

FIG. 14 is a graph showing a relationship between an energizing current of a linear solenoid and a control back pressure.

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

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

[Explanation of symbols]

 DESCRIPTION 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 / 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

──────────────────────────────────────────────────続 き Continued on the front page (58) Field surveyed (Int.Cl. ⁷ , DB name) F16H 9/00 F16H 59/00-63/48

Claims (2)

(57) [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 the drive-side and driven-side movable pulleys, and a pulley of the drive-side movable pulley. A drive-side cylinder for setting a width, a driven-side cylinder for setting a pulley width of the driven-side movable pulley, and clutch means for controlling connection / disconnection of power between the drive source and the drive-side movable pulley. Transmission torque calculating means for calculating a transmission torque transmitted from the drive-side movable pulley via the V-belt,
A side pressure control valve for adjusting a side pressure control oil pressure supplied to the drive-side and driven-side cylinders based on the transmission torque; and a clutch control means for controlling operation of the clutch means. The drive means includes: a drive torque from the drive source; and an inertia of an output rotation member of the drive source, when the clutch control unit engages the clutch unit and the power transmission is performed.
Shark torque and inertia of the drive side movable pulley
Calculating the transmission torque based on the transmission torque, and when the clutch transmission is released by the clutch control means and the power transmission is interrupted, the drive source
Wherein the transmission torque is calculated based on the inertia torque of the drive-side movable pulley without using the drive torque from the pulley. 2. The transmission torque calculation means, wherein when the clutch control means releases the clutch means and the power transmission is cut off, the control oil pressure in the drive side cylinder is controlled in the driven side cylinder. If the pressure is lower than the oil pressure, the transmission torque is calculated based only on the inertia torque of the drive-side movable pulley. If the control oil pressure in the driven-side cylinder is lower than the control oil pressure in the drive-side cylinder, the drive-side movable The pulley side pressure control device for a belt-type continuously variable transmission according to claim 1, wherein a transmission torque is calculated based on an inertia torque of a pulley and a friction torque of the V-belt.