JP7258446B2 - Shelf pressure estimator - Google Patents

Description

The present invention relates to a shelf pressure estimating device for estimating the shelf pressure of a clutch.

2. Description of the Related Art A conventional vehicle that runs by the power of an engine is equipped with a transmission. The power of the engine is input to the input shaft of the transmission, changed within the transmission, and transmitted from the output shaft of the transmission to the left and right drive wheels via a differential gear or the like.

As a transmission, a continuously variable transmission (CVT) and a stepped automatic transmission (AT) are widely known. These transmissions are equipped with a plurality of hydraulic clutches (brakes), and depending on the combination of engagement/disengagement of the plurality of clutches, P range (parking range), R range (reverse range), N range A range (neutral range) and a D range (forward range) are configured.

In the engagement control for engaging the clutch, it is important to control the transmission torque capacity in order to prevent occurrence of engagement shock. Since the transmission torque capacity of the clutch depends on the hydraulic pressure (clutch pressure) supplied to the clutch, the transmission torque capacity of the clutch can be controlled by controlling the clutch pressure. However, the relationship between the clutch pressure and the transmission torque capacity changes due to factors such as changes in the piston stroke of the clutch over time. Therefore, the relationship between the clutch pressure and the transmission torque capacity, for example, the shelf pressure, which is the clutch pressure when the transmission torque capacity of the clutch changes, is learned.

JP-A-6-159493

For example, in a configuration in which the transmission is a belt-type continuously variable transmission and a clutch is interposed between the input shaft to which power is input from the torque converter and the primary shaft, the clutch pressure is increased from 0 to The clutch pressure when a drop in the turbine speed of the torque converter is detected can be estimated as the shelf pressure. However, it is difficult to distinguish from a decrease in the turbine rotation speed due to drag torque, and the estimated shelf pressure tends to vary.

Also, in this configuration, it is conceivable to decrease the clutch pressure from the state in which the clutch is fully engaged, and to estimate the clutch pressure when an increase in the turbine speed is detected as the shelf pressure. However, since the increase in the turbine speed is affected by the torque amplification of the torque converter, the shelf pressure cannot be accurately estimated.

SUMMARY OF THE INVENTION An object of the present invention is to provide a shelf pressure estimating device capable of accurately estimating the shelf pressure of a friction engagement element.

In order to achieve the above object, the shelf pressure estimating device according to the present invention has a plate and a disk facing each other in the direction of the rotation axis. When the piston is engaged by pressing the plate against the disk and the oil pressure in the oil chamber is released from the engaged state, the piston's pressure increases with the passage of time. A device for estimating shelf pressure, which is the hydraulic pressure of the oil chamber when the transfer torque capacity of the friction engagement element switches between gains and losses, targeting friction engagement elements in which the oil pressure in the oil chamber decreases linearly until movement stops. wherein, when the hydraulic pressure in the oil chamber is released from the state in which the frictional engagement element is engaged, the time rate of change of the hydraulic pressure in the oil chamber that linearly decreases due to the elastic force of the return spring and the movement of the piston The hydraulic pressure in the oil chamber is detected at the time when the hydraulic pressure is stopped, and a value calculated according to an arithmetic expression using the detected rate of change over time and the hydraulic pressure is estimated as the shelf pressure.

Since the hydraulic pressure in the oil chamber can be detected, for example, by a hydraulic sensor, the time rate of change of the hydraulic pressure that linearly decreases due to the elastic force of the return spring can be calculated, for example, by time-differentiating the hydraulic pressure detected by the hydraulic pressure sensor. can. In addition, since the hydraulic pressure remaining in the oil chamber drops sharply when the movement of the piston stops, the hydraulic pressure in the oil chamber at the time when the movement of the piston stops can be easily determined.

Therefore, by calculating the shelf pressure of the friction engagement element according to an arithmetic expression using the time rate of change of the hydraulic pressure that linearly decreases due to the elastic force of the return spring and the hydraulic pressure of the oil chamber at the time when the movement of the piston stops, The shelf pressure can be estimated with high accuracy.

The friction engagement element has a cushioning spring disposed on the other side of the plate on the other side opposite to the one side in the rotation axis direction, and the piston moves to one side against the elastic force of the return spring. Alternatively, after the piston comes into contact with the cushioning spring, the piston presses the plate against the disk against the elastic forces of the return spring and the cushioning spring, so that the engagement is achieved.

In the case of a frictional engagement element with such a configuration, the shelf pressure is P1, the oil pressure in the oil chamber at the time when the piston stops moving is P2, and the time from the release of the oil pressure in the oil chamber to the separation of the piston from the cushioning spring. Let T1 be the time, T2 be the time from when the supply of hydraulic pressure to the oil chamber is stopped until the movement of the piston stops, and α2 be the time change rate of the hydraulic pressure that linearly decreases due to the elastic force of the return spring,
The formula is
P1=P2-α2(T2-T1)
may be

Further, let P1 be the shelf pressure, P0 be the hydraulic pressure of the oil chamber before the hydraulic pressure of the oil chamber is released, P2 be the hydraulic pressure of the oil chamber when the movement of the piston stops, and P2 be the hydraulic pressure of the oil chamber from the release of the hydraulic pressure of the oil chamber. T2 is the time until the movement stops, α1 is the time change rate of the hydraulic pressure that linearly decreases due to the elastic force of the return spring and the cushioning spring, and α1 is the time change of the hydraulic pressure that linearly decreases due to the elastic force of the return spring. Assuming the ratio α2,
The formula is
P1=(α1×P2−α2×P0−α1×α2×T2)/(α1−α2)
may be

According to the present invention, it is possible to accurately estimate the shelf pressure of the friction engagement element.

1 is a skeleton diagram showing the configuration of a vehicle drive system according to an embodiment of the present invention; FIG. FIG. 3 is a cross-sectional view showing the configuration of a forward clutch; FIG. 4 is a diagram showing the relationship between the clutch pressure of the forward clutch and the piston stroke; 2 is a block diagram showing the configuration of a vehicle control system; FIG. FIG. 4 is a diagram showing an example of temporal change in clutch pressure of a forward clutch;

BEST MODE FOR CARRYING OUT THE INVENTION Below, embodiments of the present invention will be described in detail with reference to the accompanying drawings.

<Vehicle drive system>
FIG. 1 is a skeleton diagram showing the configuration of the driving system of a vehicle 1 according to one embodiment of the invention.

A vehicle 1 is equipped with an engine 2 as a drive source, and employs, for example, a front-engine rear-wheel-drive (FR) layout. The engine 2 is mounted on the front portion of the vehicle 1 in a vertical orientation such that the crankshaft 3 is oriented vertically with respect to the front-rear direction of the vehicle 1 (hereinafter simply referred to as "the front-rear direction").

The engine 2 is, for example, a 3-cylinder 4-stroke engine, but is not limited to a 3-cylinder 4-stroke engine. That is, the number of cylinders of the engine 2 is not limited to three, and may be four or more, or may be two or less. Further, the number of strokes of the engine 2 is not limited to 4 strokes, and may be 2 strokes. The engine 2 includes an electronic throttle valve for adjusting the amount of intake air into the combustion chamber of the engine 2, an injector (fuel injection device) for injecting fuel into the intake air, and a spark plug for generating electrical discharge in the combustion chamber. are provided. A starter motor for cranking the engine 2 is attached to the engine 2 .

Power of the engine 2 is input to the transmission unit 4 . Power output from the transmission unit 4 is transmitted to a differential gear 6 via a propeller shaft 5, and from the differential gear 6 to left and right drive wheels (rear wheels) 7L and 7R.

The transmission unit 4 includes a torque converter 8 and a continuously variable transmission (CVT) 9 in a unit case forming an outer shell.

The torque converter 8 is a torque converter with a lockup mechanism, and includes a front cover 11 , a pump impeller 12 , a turbine runner 13 and a lockup clutch (lockup piston) 14 .

The front cover 11 extends in a substantially disk shape centering on a rotational axis extending in the front-rear direction, and has a shape in which an outer peripheral end thereof is bent rearward, which is the side opposite to the engine 2 side. A crankshaft 3 of the engine 2 is coupled to the center of the front cover 11 so as not to rotate relative to it.

The pump impeller 12 is arranged behind the front cover 11 . An outer peripheral end portion of the pump impeller 12 is connected to an outer peripheral end portion of the front cover 11 , and the pump impeller 12 is provided so as to be rotatable together with the front cover 11 .

Turbine runner 13 is arranged between front cover 11 and pump impeller 12 .

Lockup clutch 14 is positioned between front cover 11 and turbine runner 13 . When the hydraulic pressure in the engagement side oil chamber 15 on the turbine runner 13 side with respect to the lockup clutch 14 is higher than the hydraulic pressure in the release side oil chamber 16 on the front cover 11 side, the lockup clutch 14 is disengaged from the front cover by the differential pressure. 11 side, the lockup clutch 14 is pressed against the front cover 11, and the pump impeller 12 and the turbine runner 13 are directly connected (lockup on).

Conversely, when the hydraulic pressure in the release side oil chamber 16 is higher than the hydraulic pressure in the engagement side oil chamber 15, the lockup clutch 14 moves toward the turbine runner 13 due to the pressure difference. When the lockup clutch 14 is separated from the front cover 11, the direct connection between the pump impeller 12 and the turbine runner 13 is released (lockup off). When the engine torque rotates the pump impeller 12 in the lockup off state, oil flows from the pump impeller 12 toward the turbine runner 13 . This oil flow is received by the turbine runner 13 to rotate the turbine runner 13 . At this time, an amplifying action of the torque converter 8 occurs, and torque larger than the engine torque is generated in the turbine runner 13 .

The continuously variable transmission 9 has an input shaft 21 , a continuously variable transmission mechanism 22 , a reverse transmission mechanism 23 and an output shaft 24 . The continuously variable transmission 9 is provided so that the input shaft 21 is vertically oriented so as to extend in the front-rear direction.

The input shaft 21 extends along the rotational axis of the torque converter 8 and is provided rotatably integrally with the turbine runner 13 of the torque converter 8 . An input shaft gear 25 is integrally formed with the input shaft 21, or an input shaft gear 25 formed separately is supported so as not to be relatively rotatable.

The continuously variable transmission mechanism 22 is wound around a primary shaft 31, a secondary shaft 32, a primary pulley 33 supported by the primary shaft 31, a secondary pulley 34 supported by the secondary shaft 32, and the primary pulley 33 and the secondary pulley 34. A belt 35 is provided.

The primary shaft 31 extends parallel to the input shaft 21 with its axis spaced from the axis of the input shaft 21 in the lower right direction when viewed from the rear side. The secondary shaft 32 extends parallel to the input shaft 21 with its axis center spaced apart from the axis center of the input shaft 21 to the upper left when viewed from the rear side. In this manner, the primary shaft 31 and the secondary shaft 32 are arranged separately on the left and right with respect to the input shaft 21 . As a result, the distance between the primary shaft 31 and the secondary shaft 32 in the vertical direction can be shortened, and the size of the continuously variable transmission 9 in the vertical direction can be reduced. Therefore, even if the vehicle 1 is a vehicle such as a commercial vehicle with a low-floor cabin, the transmission unit 4 can be mounted on the vehicle 1 while ensuring the minimum ground clearance of the vehicle 1. .

The primary pulley 33 is arranged opposite to the primary fixed sheave 41 fixed to the primary shaft 31 with the belt 35 interposed therebetween, and is supported by the primary shaft 31 so as to be movable in the axial direction thereof and not relatively rotatable. and a primary movable sheave 42 . The primary movable sheave 42 is arranged on the front side with respect to the primary fixed sheave 41 .

A cylinder 43 is provided on the side opposite to the primary fixed sheave 41 side with respect to the primary movable sheave 42 , that is, on the front side. The cylinder 43 has an inner peripheral end fixed to the primary shaft 31, extends radially from the primary shaft 31, and has an outer peripheral end bent rearward. The outer peripheral end of the primary movable sheave 42 is in liquid-tight contact with the outer peripheral end of the cylinder 43 from the inner side in the radial direction of rotation. An oil chamber (piston chamber) 44 is formed between the primary movable sheave 42 and the cylinder 43 .

The secondary pulley 34 is arranged opposite to the secondary fixed sheave 45 fixed to the secondary shaft 32 with the belt 35 interposed therebetween, and is supported by the secondary shaft 32 so as to be movable in the axial direction thereof and not relatively rotatable. A secondary movable sheave 46 is provided. The secondary movable sheave 46 is arranged on the rear side with respect to the secondary fixed sheave 45, and the positional relationship between the secondary fixed sheave 45 and the secondary movable sheave 46 in the front-rear direction is the same as that of the primary fixed sheave 41 of the primary pulley 33 and the primary fixed sheave 41 of the primary pulley 33. The positional relationship with the movable sheave 42 is reversed.

A piston 47 is provided on the side opposite to the secondary fixed sheave 45 with respect to the secondary movable sheave 46 , that is, on the rear side. The piston 47 has an inner peripheral end fixed to the secondary shaft 32 and extends radially from the secondary shaft 32 . The outer peripheral end of the secondary movable sheave 46 extends rearward, and the outer peripheral end of the piston 47 is in liquid-tight contact with the outer peripheral end of the secondary movable sheave 46 from the inner side in the radial direction of rotation. An oil chamber 48 is formed between the secondary movable sheave 46 and the piston 47 .

In the continuously variable transmission mechanism 22, the oil pressure supplied to the oil chambers 44, 48 of the primary pulley 33 and the secondary pulley 34 is controlled to change the groove widths of the primary pulley 33 and the secondary pulley 34, whereby the belt is The gear ratio (the pulley ratio between the primary pulley 33 and the secondary pulley 34) is continuously and steplessly changed within a constant gear ratio range.

Specifically, when the belt gear ratio is decreased, the hydraulic pressure supplied to the oil chamber 44 of the primary pulley 33 is increased. As a result, the primary movable sheave 42 of the primary pulley 33 moves toward the primary fixed sheave 41, and the gap (groove width) between the primary fixed sheave 41 and the primary movable sheave 42 becomes smaller. Accordingly, the winding diameter of the belt 35 around the primary pulley 33 increases, and the interval (groove width) between the secondary fixed sheave 45 and the secondary movable sheave 46 of the secondary pulley 34 increases. As a result, the belt transmission ratio becomes smaller.

When the belt gear ratio is increased, the hydraulic pressure supplied to the oil chamber 44 of the primary pulley 33 is decreased. As a result, the thrust of the secondary pulley 34 against the belt 35 becomes greater than the thrust of the primary pulley 33 against the belt 35 , the distance between the secondary fixed sheave 45 and the secondary movable sheave 46 of the secondary pulley 34 becomes smaller, and the primary fixed sheave 41 becomes smaller. and the primary movable sheave 42 becomes larger. As a result, the belt transmission ratio is increased.

Although not shown, the oil chamber 48 of the secondary pulley 34 is provided with a bias spring. One end of the bias spring is in elastic contact with the secondary movable sheave 46 and the other end is in elastic contact with the piston 47 . The elastic force of the bias spring urges the secondary movable sheave 46 and the piston 47 away from each other. The secondary movable sheave 46 is applied with an urging force by the hydraulic pressure of the oil chamber 48 and a bias spring, and the belt 35 is applied with a clamping force corresponding thereto.

A primary input gear 51 is rotatably supported on the front end of the primary shaft 31 .

A forward clutch 52 is provided between the primary input gear 51 and the primary pulley 33 arranged behind it. Forward clutch 52 is a hydraulic friction clutch that is hydraulically engaged to inhibit rotation of primary input gear 51 relative to primary shaft 31 . Therefore, when the forward clutch 52 is engaged, the primary shaft 31 rotates integrally with the primary input gear 51 when the primary input gear 51 rotates. When the hydraulic pressure is released from the engaged forward clutch 52, the forward clutch 52 is released. Disengagement of the forward clutch 52 permits rotation of the primary input gear 51 with respect to the primary shaft 31 , and even if the primary input gear 51 rotates, the rotation is not transmitted to the primary shaft 31 .

A secondary input gear 53 is rotatably supported on the front end of the secondary shaft 32 .

A reverse clutch 54 is provided between the secondary input gear 53 and the secondary pulley 34 arranged behind it. The reverse clutch 54 is a hydraulic friction clutch that is hydraulically engaged to prohibit rotation of the secondary input gear 53 with respect to the secondary shaft 32 . Therefore, when the secondary input gear 53 rotates, the secondary shaft 32 rotates together with the secondary input gear 53 . When the hydraulic pressure is released from the engaged reverse clutch 54, the reverse clutch 54 is released. The disengagement of the reverse clutch 54 allows the rotation of the secondary input gear 53 with respect to the secondary shaft 32 , and even if the secondary input gear 53 rotates, the rotation is not transmitted to the secondary shaft 32 .

The reverse transmission mechanism 23 is a mechanism that transmits the power (rotation) of the input shaft 21 to the secondary shaft 32 without passing through the continuously variable transmission mechanism 22 . Reverse transmission mechanism 23 includes a reverse idler shaft 55 , a first reverse gear 56 and a second reverse gear 57 .

The reverse idler shaft 55 extends in the front-rear direction parallel to the input shaft 21 .

The first reverse gear 56 is formed integrally with the reverse idler shaft 55 or formed separately from the reverse idler shaft 55 and supported by the reverse idler shaft 55 so as not to rotate relative to it.

The output shaft 24 is arranged on the same axis as the input shaft 21 with a gap on the rear side of the input shaft 21 . An output shaft gear 58 is formed integrally with the output shaft 24, or an output shaft gear 58 formed separately from the output shaft 24 is supported so as not to rotate relative to it. Correspondingly, a secondary output gear 59 is supported on the secondary shaft 32 adjacent to the rear side of the piston 47 of the secondary pulley 34 by spline engagement so as not to rotate relative to the secondary shaft 32 . The output shaft gear 58 and the secondary output gear 59 are in mesh with each other.

The continuously variable transmission 9 has, for example, a P range (parking range), an R range (reverse range), an N range (neutral range) and a D range (forward range). A shift lever (select lever) is provided in the interior of the vehicle 1 to instruct switching of the shift range. In the range of motion of the shift lever, P position, R position, N position and D position are set corresponding to P range, R range, N range and D range, respectively.

When the shift lever is in the P position, both the forward clutch 52 and the reverse clutch 54 are released and, for example, the parking gear supported by the output shaft 24 so as not to rotate relative to each other is fixed, thereby setting the P range. Configured. When the shift lever is in the N position, both the forward clutch 52 and the reverse clutch 54 are disengaged and the parking lock gear is not locked, thereby forming the N range.

When the shift lever is in the D position, the forward clutch 52 is engaged and the reverse clutch 54 is released, thereby forming the D range. At this time, the power input from the engine 2 to the input shaft 21 via the torque converter 8 is transmitted from the input shaft gear 25 to the primary shaft 31 via the primary input gear 51 due to the engagement of the forward clutch 52 . On the other hand, even if the power input to the input shaft 21 is transmitted from the input shaft gear 25 to the secondary input gear 53 and the secondary input gear 53 rotates, the secondary input gear 53 is rotated by the secondary shaft 32 due to the release of the reverse clutch 54 . , and power is not transmitted to the secondary shaft 32 .

The power transmitted to the primary shaft 31 is changed at a belt gear ratio corresponding to the pulley ratio between the primary pulley 33 and the secondary pulley 34 and transmitted to the secondary shaft 32 . The power transmitted to the secondary shaft 32 is transmitted from the secondary output gear 59 to the output shaft 24 via the output shaft gear 58 and then transmitted from the output shaft 24 to the propeller shaft 5 .

When the shift lever is in the R position, the forward clutch 52 is disengaged and the reverse clutch 54 is engaged, thereby forming the R range. At this time, the power input from the engine 2 to the input shaft 21 via the torque converter 8 is transferred from the input shaft gear 25 to the secondary shaft 32 via the reverse transmission mechanism 23 and the secondary input gear 53 due to the engagement of the reverse clutch 54 . is transmitted to At this time, the secondary shaft 32 rotates in a direction opposite to that when the vehicle 1 moves forward. On the other hand, even if the power input to the input shaft 21 is transmitted from the input shaft gear 25 to the primary input gear 51 and the primary input gear 51 rotates, the disengagement of the forward clutch 52 causes the primary input gear 51 to rotate to the primary shaft 31 . , and power is not transmitted to the primary shaft 31 .

The power transmitted to the secondary shaft 32 is transmitted from the secondary output gear 59 to the output shaft 24 via the output shaft gear 58 and then transmitted from the output shaft 24 to the propeller shaft 5 .

<Structure of Forward Clutch>
FIG. 2 is a cross-sectional view showing the configuration of the forward clutch 52. As shown in FIG.

The forward clutch 52 has a clutch drum 61 , a clutch hub 62 and a clutch piston 63 .

The clutch drum 61 has an inner peripheral end fixed to the primary shaft 31 and extends radially from the primary shaft 31, and an outer peripheral end bent and extending toward the primary input gear 51 side, that is, toward the front side.

The clutch hub 62 is integrally formed with the primary input gear 51, has a cylindrical shape extending rearward from the primary input gear 51, and is spaced from the inner side in the axial radial direction of the outer peripheral end of the clutch drum 61. facing each other.

The clutch piston 63 is provided between the clutch drum 61 and the clutch hub 62 so as to be axially movable. The clutch piston 63 is in liquid-tight contact with the clutch drum 61 , and an oil chamber 64 is formed between the clutch drum 61 and the clutch piston 63 to supply hydraulic pressure acting on the clutch piston 63 . . Also, the clutch piston 63 is elastically biased rearward by a return spring 65 .

In the space sandwiched between the outer peripheral end of the clutch drum 61 and the clutch hub 62 in the axial radial direction, a substantially circular clutch plate 66 held by the clutch drum 61 and a substantially circular clutch disk held by the clutch hub 62 are provided. 67 are alternately arranged in that order from the rear side, which is one side in the rotation axis direction. A substantially annular cushioning spring 68 is arranged behind the rearmost clutch plate 66 . The cushioning spring 68 has its outer peripheral end held by the clutch drum 61 .

FIG. 3 is a diagram showing the relationship between clutch pressure P (oil pressure in oil chamber 64) and piston stroke (movement amount of clutch piston 63).

When hydraulic pressure is supplied to the oil chamber 64 and the clutch pressure P, which is the hydraulic pressure in the oil chamber 64 , increases, the clutch pressure P causes the clutch piston 63 to move forward against the elastic force of the return spring 65 . When the clutch pressure P rises to the contact pressure, the clutch piston 63 contacts the cushioning spring 68 . When the clutch pressure P further increases, the clutch piston 63 pushes the cushioning spring 68 forward, and the cushioning spring 68 begins to elastically deform. After that, as the clutch pressure P increases, the clutch piston 63 moves forward against the resultant force of the elastic forces of the return spring 65 and the cushioning spring 68 . The transmission torque capacity of the forward clutch 52 increases as the clutch piston 63 presses the clutch plate 66 forward through the cushioning spring 68 and the pressure contact force between the clutch plate 66 and the clutch disc 67 increases. When the cushioning spring 68 is completely pushed by the clutch piston 63 (when the elastic deformation of the cushioning spring 68 stops), the movement of the clutch piston 63 stops.

When the forward clutch 52 is engaged, rotation of the primary input gear 51 with respect to the primary shaft 31 is prohibited, and when the primary input gear 51 rotates, the primary shaft 31 rotates together with the primary input gear 51 .

When the clutch pressure P is released from the engaged state of the forward clutch 52, the resultant force of the respective elastic forces of the return spring 65 and the cushioning spring 68 causes the clutch piston 63 to move rearward, and the clutch plates 66 and the clutch discs 67 are moved. is weakened, and the transmission torque capacity of the forward clutch 52 is accordingly reduced. When the cushioning spring 68 is restored and the clutch piston 63 is separated from the cushioning spring 68, the transmission torque capacity of the forward clutch 52 becomes 0 and the forward clutch 52 is completely released.

Disengagement of the forward clutch 52 permits rotation of the primary input gear 51 with respect to the primary shaft 31 , and even if the primary input gear 51 rotates, the rotation is not transmitted to the primary shaft 31 .

<Vehicle control system>
FIG. 4 is a block diagram showing the configuration of the control system of the vehicle 1. As shown in FIG.

The vehicle 1 is equipped with an ECU (Electronic Control Unit) including a microcomputer (microcontroller unit). A microcomputer includes, for example, a CPU, a nonvolatile memory such as a flash memory, and a volatile memory such as a DRAM (Dynamic Random Access Memory). Although only one ECU 71 is shown in FIG. 2, the vehicle 1 is equipped with a plurality of ECUs having the same configuration as the ECU 71 in order to control each part. A plurality of ECUs including the ECU 71 are connected so as to be capable of two-way communication using a CAN (Controller Area Network) communication protocol.

The ECU 71 incorporates engine control logic for controlling the engine 2 and shift control logic for controlling the shift unit 4 . The engine control logic controls an electronic throttle valve, injectors, spark plugs, and the like provided in the engine 2 for starting, stopping, adjusting the output of the engine 2, and the like. Hydraulic pressure is supplied to each part of the transmission unit 4 by the transmission control logic to switch the lockup ON/OFF of the torque converter 8 and control the belt gear ratio of the continuously variable transmission mechanism 22 (the gear ratio of the continuously variable transmission 9). Various valves and the like included in the hydraulic circuit for supplying are controlled.

Various sensors required for control are connected to the ECU 71, and detection signals of the various sensors are input. Various sensors include, for example, an oil pressure sensor 72 that detects the clutch pressure P of the forward clutch 52 (the oil pressure of the oil chamber 64). The ECU 71 acquires various information from the detection signals of various sensors. Further, the ECU 71 receives various information necessary for control from other ECUs. Information input to the ECU 71 from another ECU may be obtained by connecting a sensor for obtaining the information to the ECU 71 and obtaining the information from the detection signal of the sensor in the ECU 71 .

<Shelf pressure estimation processing>
FIG. 5 is a diagram showing an example of changes in the clutch pressure P over time.

When the shift lever is operated from the D position to another position, the operation of the shift lever is transmitted through the cable to the manual valve of the automatic transmission, and the manual valve is operated, thereby shifting the forward clutch from the manual valve. Hydraulic output to engage 52 stops. Along with this, the clutch pressure P of the forward clutch 52 (the hydraulic pressure of the oil chamber 64) is released. Further, when the ECU 71 changes the clutch command pressure for the valve for regulating the hydraulic pressure supplied to the oil chamber 64 of the forward clutch 52 from the engaged state of the forward clutch 52 to 0 MPa, the clutch pressure P of the forward clutch 52 is changed to 0 MPa. To be released.

When the clutch pressure P is released (time T0), the resultant force of the respective elastic forces of the return spring 65 and the cushioning spring 68 causes the clutch piston 63 to move rearward. Along with this, the clutch pressure P linearly decreases from the oil pressure (engagement pressure) P0 at a time change rate (slope) α1 (time T0-T1).

When the clutch piston 63 is separated from the cushioning spring 68, the elastic force of the return spring 65 causes the clutch piston 63 to move rearward. Along with this, the clutch pressure P linearly decreases at a time rate of change (inclination) α2. When the clutch piston 63 comes into contact with the clutch drum 61 and stops moving (time T2), the volume of the oil chamber 64 ceases to change, and the clutch pressure P sharply drops from the clutch pressure P2.

In the ECU 71, when the clutch pressure P is released, the clutch pressure P at the time T1 when the monotonous decrease of the clutch pressure P at the time change rate α1 ends, that is, at the timing when the clutch piston 63 moves away from the cushioning spring 68. A shelf pressure P1, which is the clutch pressure P, is estimated by calculation. The shelf pressure P1 can be calculated by calculation using the clutch pressure P2 at the time T2 and the time rate of change α2 of the clutch pressure P. That is, the shelf pressure P1 can be calculated according to the following arithmetic expression (1) or (2).

P1=P2-α2(T2-T1) (1)

P1=(α1×P2−α2×P0−α1×α2×T2)/(α1−α2)
... (2)

The arithmetic expression (2) can be derived by solving the expression (3) for "T1", substituting the expression for "T1" into the expression (4), and solving the expression (4) for "P1". .
P1=P0+α1×T1 (3)
P1=P2-α2(T2-T1)=P2-α2×T2+α2×T1 (4)

<Effect>
A hydraulic pressure sensor 72 can detect the clutch pressure P0 before the clutch pressure P starts to decrease and the clutch pressure P2 at time T2 when the clutch piston 63 stops moving. As the clutch piston 63 stops moving, the clutch pressure P sharply drops, so the clutch pressure P2 at the time when the clutch piston 63 stops moving can be easily identified.

The time rate of change α1 and the time rate of change α2 of the clutch pressure P and the time T1 at which the monotonous decrease at the time rate of change α1 of the clutch pressure P is completed are calculated using the clutch pressure P detected by the oil pressure sensor 72. can. That is, the minimum value (negative value) of the values obtained by smoothing (moving average processing, low-pass filtering) the first order differential value of the clutch pressure P can be obtained as the time rate of change α1 of the clutch pressure P. can. Further, the maximum value (negative value) of the values obtained by smoothing the first-order differential value of the clutch pressure P can be obtained as the rate of change α2 of the clutch pressure P over time. The time when the second-order differential value of the clutch pressure P exceeds a predetermined value can be obtained as the time T1 when the monotonous decrease of the clutch pressure P at the time change rate α1 ends.

Therefore, the shelf pressure P1 can be accurately estimated by calculating the shelf pressure P1 according to the arithmetic expression (1) or the arithmetic expression (2).

Since the shelf pressure P1 can be accurately estimated, the transmission torque capacity of the forward clutch 52 can be accurately controlled. Therefore, it is possible to suppress the occurrence of the engagement shock of the forward clutch 52 when shifting from the N range to the D range. In addition, when excessive torque is input to the drive wheels 7L and 7R from the road surface, the forward clutch 52 can be controlled to slip before the belt 35, so-called clutch fuse control, can be performed with high accuracy. can be suppressed satisfactorily. Furthermore, since the clutch fuse control can be performed with high accuracy, the clamping pressure applied to the belt 35 can be reduced, and the fuel efficiency of the vehicle 1 can be improved.

<Modification>
Although one embodiment of the present invention has been described above, the present invention can also be implemented in other forms.

For example, although the forward clutch 52 was taken up in the above-described embodiment, the shelf pressure estimation technique according to the present invention can also be applied to the reverse clutch 54 .

Further, the transmission unit 4 is not limited to being vertically arranged behind the engine 2 so that the input shaft 21 of the CVT 5 extends in the longitudinal direction of the vehicle. The configuration may be such that the input shaft of the CVT is placed horizontally so as to extend in the left-right direction of the vehicle.

In addition, various design changes can be made to the above configuration within the scope of the matters described in the claims.

1: vehicle 52: forward clutch (friction engagement element)
54: Reverse clutch (friction engagement element)
63: Clutch piston (piston)
64: Oil chamber 65: Return spring 66: Clutch plate (plate)
67: Clutch disc (disc)
68: Cushioning spring 71: ECU (shelf pressure estimator)
72: Oil pressure sensor

Claims (3)

A plate and a disk are provided facing each other in the direction of the rotation axis, and a cushioning spring is arranged on the other side of the plate at the end of the other side opposite to the one side in the direction of the rotation axis, and the drum and the piston. The piston moves to the one side against the elastic force of the return spring due to the hydraulic pressure supplied to the oil chamber between Engagement is achieved by pressing the plate against the disc against the elastic forces of the spring and the cushioning spring , and when the oil pressure in the oil chamber is released from the engaged state, the piston moves the return spring and the cushioning spring. Due to the elastic force of the spring, it starts to move to the other side opposite to the one side in the rotation axis direction, and after the piston starts to move, after the piston is separated from the cushioning spring, the movement of the piston is caused by the drum . The friction engagement element is targeted for a friction engagement element in which the oil pressure in the oil chamber linearly decreases with the passage of time until it is blocked and stopped, and when the transfer torque capacity of the friction engagement element changes. A device for estimating shelf pressure, which is the hydraulic pressure of an oil chamber,
A time rate of change of the hydraulic pressure in the oil chamber that linearly decreases due to the elastic force of the return spring when the hydraulic pressure in the oil chamber is released from the state in which the friction engagement element is engaged , or the return spring and detecting the time rate of change of the oil pressure in the oil chamber, which linearly decreases due to the elastic force of the cushioning spring, and the oil pressure in the oil chamber at the time when the movement of the piston stops,
A shelf pressure estimating device for estimating a value calculated according to an arithmetic expression using the detected time change rate and oil pressure as the shelf pressure. Let P1 be the shelf pressure, P2 be the oil pressure in the oil chamber when the movement of the piston stops, and T1 be the time from the release of the oil pressure in the oil chamber to the separation of the piston from the cushioning spring, Let T2 be the time from when the hydraulic pressure supply to the oil chamber is stopped until the movement of the piston stops, and α2 be the time rate of change of the hydraulic pressure that linearly decreases due to the elastic force of the return spring,
The above formula is
P1=P2-α2(T2-T1)
The shelf pressure estimating device according to claim 1, wherein: Let P1 be the shelf pressure, P0 be the oil pressure in the oil chamber before the oil pressure in the oil chamber is released, P2 be the oil pressure in the oil chamber when the movement of the piston stops, and P2 be the oil pressure in the oil chamber. Let T2 be the time from release of the hydraulic pressure to the stop of the movement of the piston, α1 be the rate of change of the hydraulic pressure that linearly decreases due to the elastic forces of the return spring and the cushioning spring, and α1 be the rate of change with time. Let α2 be the time rate of change of the hydraulic pressure that decreases linearly,
The above formula is
P1=(α1×P2−α2×P0−α1×α2×T2)/(α1−α2)
The shelf pressure estimating device according to claim 1, wherein:

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