JP2001165293A - Transmission control device for belt-type continuously variable transmission - Google Patents

Abstract

(57) Abstract: A rotary inertia force can be reduced without increasing the size of a cylinder for adjusting a groove width interval of a pulley. SOLUTION: A hydraulic oil discharged from a pump 47 driven by an engine is regulated and supplied to a secondary oil chamber 21 of a secondary side cylinder provided on a secondary shaft and changing a pulley groove width of a secondary pulley, Primary oil chamber 1 of primary side cylinder that changes the pulley groove width of primary pulley provided on primary shaft
8 is supplied with hydraulic oil from a bidirectional pump 82 driven by a motor 83 using the hydraulic pressure supplied to the secondary oil chamber 21 as a source pressure. By rotating the motor 83 in both the forward and reverse directions, the hydraulic oil from the pump 47 moves back and forth between the oil chambers 18 and 21 to change the pulley groove width, thereby controlling the gear shift.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention particularly relates to a shift control device for a belt-type continuously variable transmission which is suitable for use in shift control of a mobile machine such as an automobile.

[0002]

2. Description of the Related Art A belt-type continuously variable transmission (CV) for automobiles.
As T), a metal pulley is provided between a primary pulley having a variable pulley groove width provided on a primary shaft on a driving side and a secondary pulley having a variable pulley groove width provided on a secondary shaft on a driven side, that is, a driven side. There is a type in which a drive belt is stretched, the pulley diameters of a primary pulley and a secondary pulley are changed by hydraulic pressure, and the rotational speed of a secondary shaft is changed steplessly.

[0003] The shift control of the CVT is performed by controlling the hydraulic pressure to hydraulic cylinders provided on a primary side and a secondary side, respectively. The hydraulic pressure applied to a secondary pulley is generated by a pump driven by an engine. There is a system in which the secondary pressure applied to the secondary hydraulic cylinder is reduced to adjust the primary pressure applied to the primary hydraulic cylinder. This method is also called a master-slave method because the primary pressure does not exceed the base secondary pressure, that is, the master pressure. In order to upshift the gear ratio to higher speeds, it is necessary to pinch the belt with a stronger force on the primary pulley than with the secondary pulley, but it is not possible to pinch the primary pulley with a stronger force using only hydraulic pressure. The pressure receiving area of the primary side hydraulic cylinder is set larger than that of the secondary side hydraulic cylinder, and the pressure receiving area of the primary side is normally set to about twice that of the secondary side.

[0004] As a speed change adjusting valve for adjusting the primary pressure, for example, as described in Japanese Patent Application Laid-Open No. 57-161360, the speed of the oil supplied to the hydraulic cylinder on the primary side is controlled. In this type, the primary pressure is connected to the secondary pressure by adjusting the stroke of the valve shaft to change the opening area of each port, or the primary pressure is drained to change the speed. Control is performed. When adjusting the opening area of the speed change control valve, if the primary pressure input / output is closed when the target speed ratio is reached, the pulley ratio can be maintained. There is an advantage that the ratio can be kept stable.

However, when this flow control type shift control valve is used, it is difficult to adjust the opening area of the shift control valve, and the relationship between the flow rate and the stroke of the valve shaft must be set with high accuracy. In addition, there is a problem that the pressure receiving area of the hydraulic cylinder on the primary side is increased to increase the size of the cylinder, thereby increasing the inertia, that is, the inertia force. In order to suppress the increase in cylinder inertia,
For example, as described in JP-A-63-19742,
The double cylinder has a problem that the structure is complicated.

[0006] As a shift adjusting valve for adjusting the primary pressure, for example, there is a pressure adjusting type using a pressure reducing valve as described in Japanese Patent Application Laid-Open No. 3-229058. In this type, as described above, there is a problem that the cylinder becomes large and inertia increases, and when a disturbance such as torque fluctuation acts, the primary pressure is changed to maintain the pulley ratio. Therefore, the pulley ratio may fluctuate without the control response of the primary pressure normally catching up with the fluctuation, and it is necessary to prevent the fluctuation.

On the other hand, a push-pull type cylinder in which the pressure receiving areas of the primary side cylinder and the secondary side cylinder are made the same has been developed.
As described in the publication, the operation of the pulley adjusts the pressure required by the larger pressure required by the primary and the secondary, and supplies the pressure to a cylinder that requires one of the higher pressures,
Cylinders requiring low pressure are supplied via a stroke adjusting valve. In this case, although an increase in the size and complexity of the hydraulic cylinder on the primary side can be avoided, both the hydraulic pressures in the primary and secondary cylinders must be controlled using a shift adjustment valve, which complicates the hydraulic circuit. There is a problem.

[0008]

As described above, in the master-slave system, the cylinder area on the primary side needs to be larger than that on the secondary side. Therefore, the cylinder becomes large, its inertia increases, and the acceleration performance is impaired. Will be. In addition, a double cylinder structure complicates the mechanism. Further, in the push-pull type, the cylinder on the primary side can be reduced in size, but it is necessary to provide a shift adjusting valve on both the primary side and the secondary side.

An object of the present invention is to make it possible to reduce the rotational inertia force of a pulley without increasing the size of a cylinder for adjusting the groove width interval of the pulley.

[0010]

According to the present invention, there is provided a shift control device for a belt-type continuously variable transmission, which comprises a primary pulley provided on a primary shaft driven by an engine and a primary pulley provided on a secondary shaft. A shift control device for a belt-type continuously variable transmission having a secondary pulley around which a drive belt is stretched, wherein the primary cylinder changes a pulley groove width of the primary pulley, and the secondary cylinder changes a pulley groove width of the secondary pulley. A cylinder, an oil pump driven by the engine to regulate a discharge flow and supply the same to one of the cylinders, and a positive displacement fluid machine driven in both forward and reverse directions by a motor to supply the discharge flow to the other cylinder And characterized in that:

In the transmission control apparatus for a belt-type continuously variable transmission according to the present invention, a drive belt is stretched between a primary pulley provided on a primary shaft driven by an engine and the primary pulley provided on a secondary shaft. A shift control device for a belt-type continuously variable transmission having a secondary pulley, the primary cylinder changing a pulley groove width of the primary pulley, a secondary cylinder changing a pulley groove width of the secondary pulley, and the engine An oil pump that is driven and regulates the discharge flow by a pressure regulating valve and supplies it to one of the cylinders, and is driven in both forward and reverse directions by a motor, and discharges the discharge flow to the other cylinder using the regulated hydraulic pressure as a source pressure. A two-way pump for supplying the two-way pump and a reversible rotary motor for rotating the two-way pump in both forward and reverse directions. And having a motor.

In the speed change control device for a belt-type continuously variable transmission according to the present invention, a drive belt is stretched between a primary pulley provided on a primary shaft driven by an engine and the primary pulley provided on a secondary shaft. A shift control device for a belt-type continuously variable transmission having a secondary pulley, the primary cylinder changing a pulley groove width of the primary pulley, a secondary cylinder changing a pulley groove width of the secondary pulley, and the engine An oil pump that is driven and supplies a discharge flow to one of the cylinders; a pressure adjustment control unit that adjusts a hydraulic pressure of hydraulic oil supplied to the one cylinder in accordance with a traveling state of the vehicle; And supplies the discharge flow to the other cylinder using the regulated hydraulic pressure as the original pressure. A bidirectional pump, a reversible rotary motor that drives the bidirectional pump in both forward and reverse directions, and a motor control unit that controls the reversible motor in accordance with a running state such as an engine speed and a throttle opening. Features.

According to the present invention, a high-pressure oil pump is provided in the high-pressure oil passage connected to each of the cylinders, and the low-pressure oil passage is provided separately from the high-pressure oil passage. May be provided.

[0014]

Embodiments of the present invention will be described below in detail with reference to the drawings.

FIG. 1 is a schematic view showing an example of a drive system of a belt-type continuously variable transmission, that is, a CVT. A crankshaft 1 driven by an engine (not shown) includes a front cover 3 of a pump side case 3 of a torque converter 2. 3a is directly connected to the pump side case 3 via a drive plate 4.
A turbine runner 5 disposed opposite to a pump impeller 3 b provided therein is directly connected to a turbine shaft 6.
A stator 7 is arranged between the pump impeller 3b and the turbine runner 5, and the stator 7 is supported by a one-way clutch 8a attached to a stator support shaft 8. A lock-up clutch 9 is directly connected to the turbine shaft 6 so as to be movable between an engagement position for engaging the drive plate 4 and an open position away from the drive plate 4. Engine power is transmitted via the torque converter 2 or the lock-up clutch 9. The power is transmitted to the turbine shaft 6.

One side of the lock-up clutch 9 is a supply chamber, that is, an apply chamber 9a, and the other side is an open chamber, that is, a release chamber 9b. The hydraulic pressure supplied to the release chamber 9b is circulated through the apply chamber 9a. Thus, the torque converter 2 enters an operating state. On the other hand, by supplying hydraulic pressure to the apply chamber 9a and reducing hydraulic pressure in the release chamber 9b, the lock-up clutch 9 is engaged with the front cover 3a to enter a lock-up state. By adjusting the pressure in the release chamber 9b, slip pressure control is performed so that the lock-up clutch 9 is slid.

The turbine shaft 6 is transmitted to an input shaft of a continuously variable transmission 12, that is, a primary shaft 13 via a forward / reverse switching device 11. A primary pulley 14 is provided on the primary shaft 13, and the primary pulley 14 is a fixed pulley 14 a fixed to the primary shaft 13.
And a movable pulley 14b opposed to the primary shaft 13 so as to be freely slidable in the axial direction by a ball spline or the like. The cone surface distance of the pulley, that is, the pulley groove width is variable. . An output shaft arranged in parallel with the primary shaft 13, that is, a secondary shaft 15 is provided with a secondary pulley 16.
6 is a fixed pulley 16a fixed to the secondary shaft 15
And a movable pulley 16b opposed to the secondary shaft 15 so as to be slidable in the axial direction on the secondary shaft 15 similarly to the movable pulley 14b, and the groove width of the pulley is variable. Note that the entire drive system is incorporated in the case 10.

A drive belt 17 is stretched between the primary pulley 14 and the secondary pulley 16, and by changing the groove width of both pulleys 14, 16,
By changing the ratio of the winding diameter to each of the pulleys 14 and 16, the rotational speed of the secondary shaft 15 is continuously variable.

In order to change the groove width of the primary pulley 14, a primary oil chamber 18 is provided between the primary pulley 14 and the movable pulley 14b.
Is mounted on the primary shaft 13, and a plunger 22 forming a secondary oil chamber 21 with the movable pulley 16 b is mounted on the secondary shaft 15 in order to change the groove width of the secondary pulley 16. The plunger 22 forms a cylinder on the secondary side.

The secondary shaft 15 is connected to an intermediate shaft 25 via gears 23 and 24. A gear 26 attached to the intermediate shaft 25 meshes with a final gear 28 of a differential device 27, and the differential device 27
Axles 29a, 29b connected to the wheels 31a, 31
b is attached. In the case of a front wheel drive vehicle, the wheels 31a and 31b are front wheels.

The forward / reverse switching device 11 has a clutch cylinder 32 provided on a forward clutch drum fixed to the turbine shaft 6 and a clutch hub 33 fixed to the primary shaft 13. Is provided with a multi-plate type forward clutch 34. This forward clutch 34
Is incorporated in the clutch cylinder 32. Accordingly, when the hydraulic pressure is supplied to the clutch oil chamber 32 a in the clutch cylinder 32 and the forward clutch 34 is connected, the rotation of the turbine shaft 6 is transmitted to the primary shaft 13 via the clutch hub 33 and the primary shaft 13 It rotates in the same forward direction as the turbine shaft 6.

A sun gear 36 is fixed to the primary shaft 13, and a ring gear 37 is rotatably provided inside the case 10 outside the sun gear 36. Carrier 3 attached to a forward clutch drum unit having a clutch cylinder 32
A pair of planetary pinion gears 41 and 42 that mesh with each other are rotatably mounted on 8, and one planetary pinion gear 41 is meshed with the sun gear 36, and the other planetary pinion gear 42 is meshed with the internal teeth of the ring gear 37. These gears constitute a double pinion type planetary gear. The planetary pinion gears 41 and 42 are shown one by one in FIG. 1 for convenience of drawing, but mesh with each other in pairs, and a plurality of pairs are provided.

A multi-plate reverse brake 43 is provided between the ring gear 37 and the case 10, and a hydraulic piston 44 for operating the reverse brake 43 is provided.
Are incorporated in a brake cylinder 45 formed in the case 10. Accordingly, when the hydraulic pressure is supplied to the oil chamber 45a in the brake cylinder 45 to bring the reverse brake 43 into a braking state while the forward clutch 34 is released, the ring gear 37 is fixed to the case 10. The rotation of the carrier 38 together with the turbine shaft 6 causes the sun gear 36 and the primary shaft 1 via the paired planetary pinion gears 41 and 42.
3 rotates in the reverse rotation direction opposite to the turbine shaft 6. The forward clutch 34 and the reverse brake 43 are friction engagement elements of the forward / reverse switching device 11.

In order to operate hydraulic operating devices such as the brake cylinder 45 and the clutch cylinder 32, the case 1
An oil pump 47 serving as a hydraulic pressure source is disposed in the oil pump 0, and the oil pump 47 is driven by the engine by the crankshaft 1 via the pump side case 3.

FIG. 2 is a diagram showing a hydraulic circuit of a shift control device according to an embodiment of the present invention. The discharge port of an oil pump 47 that sucks and discharges oil in an oil pan 48 has a secondary pressure path, The line pressure path 50 is connected to the secondary oil chamber 21 that operates the movable pulley 16 b of the secondary pulley 16, and is also connected to the secondary pressure port of the secondary pressure adjustment valve 51. The secondary pressure supplied to the secondary oil chamber 21 is adjusted to a predetermined pressure by the secondary pressure adjusting valve 51 so that the driving belt 1
7 is controlled to a value commensurate with the required transmission capacity. That is, when the engine output is large, such as when climbing a hill or suddenly accelerating, the secondary pressure is increased to prevent the drive belt 17 from slipping. When the engine output is small, the secondary pressure is reduced, and the loss and transmission efficiency of the oil pump 47 are reduced. Is improved.

A lubricating pressure passage 52 is connected to a drain port of the secondary pressure adjusting valve 51. The pressure in the lubricating pressure passage 52 is regulated by a lubricating pressure adjusting valve 53, and a lubricating portion of the forward / reverse switching device 11 and the drive belt 17 is provided. The air is supplied to the balance chamber of the secondary shaft via the check valve 54, and is also supplied to the lock-up release chamber 9b and the like.

A clutch pressure passage 56 is connected to the line pressure passage 50 via a clutch pressure regulating valve 55 which is a pressure reducing valve. When an external pilot pressure is supplied to the clutch pressure regulating valve 55, the clutch pressure passage 56 Is set to a low pressure, and when the supply of the external pilot pressure is stopped, the pressure is set to a higher pressure than when it was supplied.

The apply pressure path 57 connected to the apply chamber 9a of the torque converter 2, the release pressure path 58 connected to the release chamber 9b, and the switching pressure for brake connected to the brake oil chamber 45a for operating the reverse brake 43 Road 5
9 and a switch valve 61 for controlling the connection between the clutch switching pressure path 60 connected to the clutch oil chamber 32a for operating the forward clutch 34, the lubrication pressure path 52 and the clutch pressure path 56, and the like. Is provided.

The switch valve 61 has four portions each having a three-port switching valve structure. As shown in FIG. 2, the F & R mode when the external pilot pressure is not applied, that is, the vehicle speed becomes lower than a predetermined value. The lock-up clutch 9 operates in two positions: an open position of the lock-up clutch 9 in the state, and an engagement position of the lock-up clutch 9 in a state in which the external pilot pressure is applied.

In the open position, the lubricating pressure path 52 and the release pressure path 58 are in communication with each other by the switch valve 61, and the cooling path 63 provided with the oil cooler 62 and the apply pressure path 57 are in communication. . As a result, the hydraulic circuit operates the torque converter 2 and enters a mode in which the hydraulic control of the forward / reverse switching device 11 can be performed, that is, the F & R mode.
b, is discharged from the apply chamber 9a, and returns to the oil pan via the oil cooler 62.

On the other hand, at the engagement position, the clutch pressure path 56 and the apply pressure path 57 are in communication with each other, and the operating oil set to the clutch pressure is supplied to the apply chamber 9a.
At this time, the slip pressure passage 64 connected to the clutch pressure passage 56 is communicated with the release pressure passage 58. A slip pressure adjusting valve 65 is provided in the slip pressure passage 64, and the slip pressure adjusting valve 65 controls the slip pressure supplied to the slip pressure passage 64 in accordance with the external pilot pressure supplied to the external pilot chamber. , The pressure is controlled to an arbitrary pressure in the range from the same pressure as the clutch pressure to zero pressure. Therefore, when the slip pressure becomes 0, the lock-up clutch 9 is engaged to enter the lock-up mode, and when the slip pressure becomes equal to the clutch pressure, the lock-up clutch 9 is released. By appropriately controlling the slip pressure, slip control of the lock-up clutch 9 for controlling the rotation difference of the lock-up clutch 9 to be constant can be performed. When the switch valve 61 is in the engaged position, the lubricating pressure passage 52
Communicates with the cooling passage 63 via the switch valve 61 to cool the hydraulic oil.

In order to supply the external pilot pressure to the slip pressure regulating valve 65, a pilot pressure passage 6 is provided between the pilot port of the slip pressure regulating valve 65 and the clutch pressure passage 56.
The pilot pressure passage 66 is provided with a pilot pressure adjusting valve 67 for controlling the pilot pressure.

A manual valve 72 and a reverse signal valve 73 are connected to a control lever, ie, a select lever 71 for switching the driving mode, provided in the passenger compartment.
Reference numeral 73 operates at five positions corresponding to the P (parking) range, R (reverse) range, N (neutral) range, D (drive) range, and Ds (sports drive) range set by the select lever 71.

The pilot pressure line 74 for connecting the clutch pressure line 56 to the external pilot chamber of the switch valve 61 through the reverse signal valve 73 is provided with a three-port solenoid type switching valve 75. When the solenoid 75a of the switching valve 75 is energized, the switch valve 61 is in the lock-up control position, that is, the engagement position of the lock-up clutch. When the energization to the solenoid 75a is turned off, the switch valve 61 is in the F & R mode position as shown in FIG. The pilot pressure path 74 is connected to an external pilot chamber of the clutch pressure adjusting valve 55 as shown by a broken line, and when the reverse signal valve 73 is set to any of the N position, the D position and the Ds position, The clutch pressure is supplied to the external pilot chamber of the clutch pressure adjusting valve 55, and the clutch pressure is set to a low pressure. On the other hand, when the reverse signal valve 73 is set to the P position and the R position other than those described above, the hydraulic pressure is not supplied to the external pilot chamber of the clutch pressure adjusting valve 55, and the clutch pressure becomes higher than the above. Is set.

A common switching pressure passage 76 is provided between the switch valve 61 and the manual valve 72. The switching pressure passage 76 communicates with the slip pressure passage 64 when the switch valve 61 is in the F & R mode position. When the switch valve 61 is at the lock-up control position, it communicates with the clutch pressure path 56. When the manual valve 72 is set to either the D range or the Ds range by operating the select lever 71, the switching pressure path 76 is in communication with the clutch switching pressure path 60 via the manual valve 72, and is switched to the R range. When it is set, it is in a state of communication with the brake switching pressure path 59.

The oil chamber 18 of the primary cylinder 19 is connected to a line pressure line 50 by a primary pressure line 81. The primary pressure line 81 is provided with a two-way pump 82 for speed change. It is designed to be driven by an electric motor 83 which is driven to rotate in both directions. The bidirectional pump 82 adjusts the flow rate of the hydraulic pressure supplied to the oil chamber 18 using the line pressure as the original pressure. A relief valve 84 is connected to the primary pressure passage 81 on the downstream side of the pump 82, and when the flow rate to the oil chamber 18 becomes excessive and the primary pressure rises above a predetermined value, the oil pan 48 This prevents the primary cylinder 19 from being returned and damaged. When the primary pressure is lower than the lubrication pressure, the lubrication pressure is guided to the primary oil chamber 18 via the lubrication pressure path 52a connected between the primary oil chamber 18 and the lubrication pressure path 52. A one-way valve 84a is provided in the lubricating pressure passage 52a. The one-way valve 84a functions as a so-called refill valve, prevents air from being mixed into the cylinder 19, and stabilizes gear shifting. Has an action.

FIG. 3 is a view showing an example of the bidirectional pump 82. This pump 82 has a housing 87 provided with a driving gear 85 constituted by an external gear and a driven gear 86 meshing with the driving gear 85. I have. The drive shaft 85a of the drive gear 85 is spline-coupled to the shaft 83a of the electric motor 83, and is driven to rotate by the motor 83. The driven gear 86 is rotatably mounted on the housing 87 by the driven shaft 86a. The rotation is driven by the engagement with the gear 85.

One port 88 of the pump 82 is connected to the oil chamber 18 by the primary pressure line 81, and the other port 8
9 is connected to the discharge port of the oil pump 47 by a primary pressure passage 81. When the motor 83 is rotated in the forward direction indicated by the arrow in FIG. 3B, the working oil is supplied to the oil chamber 18, so that the pulley width of the primary pulley 14 is reduced and the pulley ratio is increased toward the overdrive side. shift. On the other hand, when the motor 83 is rotated in the reverse direction, the operating oil in the primary oil chamber 18 is discharged to the secondary oil chamber side, so that the hydraulic oil is downshifted to the LOW side.

The bidirectional pump is not limited to the circumscribed gear pump shown in FIG. 3, but may be a positive displacement type fluid machine such as an inscribed gear pump, a vane pump or a piston pump. However, since forward rotation and reverse rotation are performed, it is preferable to use an internal or external gear pump.

FIG. 4 is a block diagram showing a control unit (TCU) 91 for controlling the operation of the shift control device shown in FIG. 2. The control unit 91 changes the running state based on signals such as a vehicle speed signal and a range signal. By making a judgment, a control signal is sent to the secondary pressure adjusting valve 51 shown in FIG.
3 to control the shift. Further, a device provided in a low-pressure system oil passage that supplies the hydraulic oil adjusted by the clutch pressure adjusting valve 55 to the clutch oil chamber 32a, the brake oil chamber 45a, and the torque converter 2 of the forward / reverse switching device 11, for example, a pilot Pressure regulating valve 67 and switching valve 7
5 is also controlled by a control signal from the control unit 91, but the details are omitted.

FIG. 5 is a block diagram showing a control logic for controlling the operation of the motor 83 in the control unit 91 shown in FIG. 4, and includes a target gear ratio calculator 92, a current gear ratio calculator 93, and a motor control. And a portion 94. The current gear ratio calculator 93 calculates the current gear ratio io based on the signals of the rotation speed Np of the primary pulley 14 and the rotation speed Ns of the secondary pulley 16. The target gear ratio calculator 92 calculates the engine speed Ne, the throttle opening, the coolant temperature of the engine coolant, the oil temperature of the hydraulic oil, that is, the oil temperature of the transmission, and the brake switch signal, in addition to the vehicle speed signal and the range signal. Calculate the target gear ratio im.

The motor control unit 94 compares the current speed ratio io with the target speed ratio im and determines whether the current speed ratio io is + or-with respect to the target speed ratio im. The motor includes a comparison unit 95 for calculating a deviation from the current gear ratio, and a motor operation state calculation unit 96 for calculating whether to rotate the motor 83 forward, reverse, or stop based on the comparison result. The signal of the operating state calculation section 96 is sent to the motor driver section 97 to control the operation of the motor 83.

FIG. 6 is a block diagram showing the details of the motor control unit 94. The motor control unit 94 has a pulley stroke conversion unit 101. The conversion unit 101 converts the target speed ratio im of the pulley ratio and the pulley ratio. Is converted into a deviation (ΔS = Sm−So) of the movable pulley 14b of the primary pulley 14 from the current speed ratio io (Δi = im−io).
This conversion is performed using a conversion formula or map data stored in the memory.

The output signals Δi and ΔS of the pulley stroke conversion unit 101 are sent to a shift speed setting unit 102, where a target shift speed di / dt is obtained, and a target pulley displacement speed dS / dt is calculated from this value. You. The target shift speed is set by the product of the deviation of the pulley ratio and the coefficient K. That is,
(Di / dt) = K · Δi, and the coefficient K is searched as a map based on parameters indicating the running state, that is, engine speed, throttle opening, vehicle speed, brake switch signal, range signal, and the like. The target pulley displacement speed is (dS / dt) = (di / dt) · (ΔS /
Δi).

The pulley displacement speed is corrected by the maximum displacement speed limiter 103 to be within the maximum and minimum values. That is, the calculated pulley displacement speed (dS / dt)
Is greater than or equal to its maximum value (dS / dt) _max , the value is set to the maximum value, and if it is less than the maximum value, the calculated pulley displacement speed is used as the pulley displacement speed. You.

The signal from the maximum displacement speed limiter 103 is sent to the target pump speed setting unit 104. Here, the target control is performed by first multiplying the target pulley displacement speed (dS / dt) by the cylinder area Ap. The flow rate Q is determined, and then the control flow rate is divided by the theoretical discharge rate Vth of the pump to determine the theoretical rotation speed Np of the pump.

Since there is actually a volumetric efficiency of the pump, in order to calculate the target pump rotational speed Nm by correcting the theoretical rotational speed Np, the pump rotational speed correcting unit 105 obtains the volumetric efficiency ηv of the pump. The rotational speed Nm is calculated. The calculation of the volumetric efficiency ηv is performed based on the values of the oil temperature Toil of the working oil, the theoretical rotation speed Np, the line pressure Ps, and the calculated primary pressure Pp. Here, the calculated primary pressure Pp is determined by the target gear ratio im, the engine speed Ne, and the throttle opening T.
It is determined by searching map data based on the values of h and the line pressure Ps.

Based on the result of the correction by the pump speed corrector 105, the target pump speed Nm and the target excitation current Im are sent from the control value setting unit 106 to the motor driver 97. The target excitation current Im is obtained from the current-torque characteristics of the motor 83 used, that is, the rotation and holding characteristics, and the driving torque is calculated based on the calculated primary pressure Pp and the line pressure Pp.
It can be calculated as a function of s and oil temperature Toil.

When the target value is equal to the current value, the motor 83 is instructed to stop. However, when the motor 83 is stopped in a state where the primary pressure and the secondary pressure are different, the motor 83 is stopped. It is necessary to apply a stop torque to the motor 83 because the pump 82 tries to rotate by hydraulic pressure. For this reason, it is preferable to use a stepping motor having a holding torque function as the motor 83.

FIG. 7A is a block diagram showing details of the motor driver 97a when the motor 83 is a stepping motor 83a.
a includes a phase control unit 107 for obtaining a drive timing for each phase winding based on the target pump speed Nm determined by the control value setting unit 106, and a driver element for each winding at the obtained timing. The driving current Im
To drive the motor 83.

FIG. 7B is a block diagram showing details of the motor driver 97b when a DC servo motor or an AC servo motor 83b is used as the motor 83. The servo motor 83b is a stepping motor 83
There is an advantage that a larger output can be obtained than a. In this case, a feedback loop is incorporated in the motor driver 97b. To this end, in these servo motors 83b, the motor is provided with a rotation angle sensor 110 and a tachometer, that is, a rotation speed sensor 111. The rotation angle detection value ΔR is sent to a multiplier 112, and the rotation speed detection value N ₀ is 113.

[0052] signals of the respective multipliers 112 and 113 is sent to the discriminator 114, when the rotational speed target value N _m is not 0 (N _m ≠ 0), the detection of the rotational speed sensor 111 by the discriminator 114 The signal of the multiplier 113 is sent to the adder 115 so that the current feedback is performed by comparing the value N ₀ with the target value N _m and the motor 8
3b is driven. On the other hand, when the target value N _m is 0 (N _m =
In 0), the operation state of the motor 83b is made to coincide with the target value by sending the signal of the multiplier 112 to the adder 115 based on the detection value ΔR of the rotation angle sensor 110.

In the above-described transmission control apparatus for a belt-type continuously variable transmission, the primary pressure is generated by the bidirectional pump 82 using the secondary pressure supplied into the secondary oil chamber 21 as the original pressure. The shift adjustment can be performed without increasing the size of the primary cylinder 19. Thus, the rotational inertia of the primary shaft can be reduced, and the acceleration performance of the vehicle can be improved. In addition, a configuration such as a double cylinder is not required, and the cylinder structure can be simplified.

Since the flow rate of the pump 82 can be almost uniquely determined by the number of revolutions, the flow rate of hydraulic oil flowing into and out of the primary oil chamber 18 can be easily controlled with high accuracy. In particular, the linearity is improved and the shift stability is improved as compared with the case where the primary pressure is adjusted by a valve.

Since the hydraulic oil only flows between the primary oil chamber 18 and the secondary oil chamber 21, the amount of oil to be supplied from the outside is extremely small, and the discharge amount of the pump 47 driven by the engine is reduced. Can be set smaller. Thereby, power loss is suppressed, heat generation of the hydraulic oil is reduced, and the oil cooler 62 can be downsized.

Even if the motor 83 stops operating due to disconnection or the like and the pump 82 stops, the primary pressure gradually changes to the secondary pressure, and the pressure receiving areas on the primary side and the secondary side are set to be substantially the same. In this case, the pulley ratio stabilizes at around 1.0. This prevents the occurrence of a phenomenon in which the speed is shifted to the LOW side to cause the engine to blow up, or the speed is shifted to the overdrive side to cause a shortage of the driving force, which is also preferable as a failure mode.

FIG. 8 is a diagram showing a hydraulic circuit of a shift control device according to another embodiment of the present invention, in which members common to those of the hydraulic circuit shown in FIG. 2 are denoted by the same reference numerals.

The shift control device shown in FIG. 8 has a low-pressure oil pump 49 driven by the engine in addition to the high-pressure pump 47 driven by the engine, similarly to the oil pump 47 described above. The discharge port of the oil pump 49 is connected to the clutch pressure path 56. Accordingly, the hydraulic oil discharged from the high-pressure side pump 47 is regulated by the secondary pressure regulating valve 51 so that the secondary oil chamber 21 of the secondary cylinder and the shift bidirectional pump 82
Is supplied to the primary oil chamber 18. On the other hand, the hydraulic oil discharged from the pump 49 on the low pressure side is regulated by the clutch pressure regulating valve 55 and passes through the low pressure system oil passage to the clutch oil chamber 32a, the brake oil chamber 45a and the torque converter 2 of the forward / reverse switching device 11. Is supplied to the apply chamber 9a and the release chamber 9b.

In this transmission control device, the primary pressure line 81, which is a high-pressure oil passage, is separated from the clutch pressure line 56, which is a low-pressure oil passage. The hydraulic oil can be moved only to the primary oil chamber 18 and the secondary oil chamber 21. As a result, the discharge amount of the high-pressure side pump 47 is sufficient to cover the leakage of the hydraulic system returning to the oil pan 48 and the replenishment generated due to the pulley stroke being non-linear with respect to the pulley ratio. Can be significantly reduced as compared with the case described above.

Further, the pump loss when one oil pump 47 is provided as shown in FIG. 2 and the sum of the pump loss when two oil pumps 47 and 49 on the high and low pressure sides are provided as shown in FIG. Comparing pump losses, the case shown in FIG. 8 is smaller as described below.

That is, in general, the driving torque T of the oil pump is represented by T = P · D / 2π. Here, the symbol D is the theoretical discharge amount of the pump, and in the case shown in FIG. 2, since the total discharge amount is once adjusted to the line pressure Ps, the pump loss T1 becomes T1 = Ps.D / 2π.

On the other hand, in the case shown in FIG. 8, the theoretical discharge amount D of the pump is D = DL + DH (DH is the high pressure side, DL
Is low pressure side), high pressure system is Ps and low pressure system is Pc
, The pump loss T2 is expressed by the following equation.

T 2 = (DH · Ps + DL · Pc) / 2π
= (D · Pc + DH (Ps−Pc)) / 2π Then, when the reduction of the loss torque is evaluated by the ratio T2 / T1, T2 / T1 = Pc / Ps + DH / D · (1-Pc / Ps)
) = DH / D + Pc / Ps (1-DH / D).

If DH / D <1 and Pc / Ps <1,
Since T2 / T1 <1, the oil pump loss is shown in FIG.
Can be reduced.

The present invention is not limited to the above-described embodiment, and it goes without saying that various modifications can be made without departing from the scope of the invention. For example, the drive system of the belt-type continuously variable transmission is not limited to the case shown in FIG. 1, and the present invention can be applied to various types such as a type having no torque converter 2. In the illustrated case, the hydraulic oil discharged from the pump 47 is supplied to the secondary oil chamber 21 and is supplied to the primary oil chamber 18 by using the bidirectional pump 82 as the original pressure. Is supplied to the primary oil chamber 18, and the oil is supplied to the primary oil chamber 18, and the oil is supplied to the primary oil chamber 18.
2, it may be supplied to the secondary oil chamber 21.

[0066]

According to the present invention, the cylinder for adjusting the groove width of the pulley can be reduced in size, so that its rotational inertia can be reduced and the acceleration performance can be improved. Further, the shift accuracy can be improved, and the shift stability can be improved. Further, the discharge amount of the pump can be set small, so that the power loss can be reduced and the heat generation of the oil can be reduced. A high-pressure side pump for supplying only to the primary and secondary side oil chambers and a low-pressure side pump for supplying hydraulic oil to a low-pressure oil passage connected to a clutch of a forward / reverse switching device are provided. Then, it is possible to reduce the discharge amount of the high-pressure side pump and achieve downsizing of the pump,
Further, pump loss can be reduced.

[Brief description of the drawings]

FIG. 1 is a schematic diagram illustrating an example of a drive system of a belt-type continuously variable transmission.

FIG. 2 is a hydraulic circuit showing a shift control device according to an embodiment of the present invention.

3A is a longitudinal sectional view showing a specific example of the bidirectional pump shown in FIG. 2, and FIG. 3B is a transverse sectional view of FIG.

FIG. 4 is a block diagram illustrating a control unit that controls the operation of the shift control device illustrated in FIG. 2;

FIG. 5 is a block diagram showing control logic for controlling operation of a motor in the control unit shown in FIG. 4;

FIG. 6 is a block diagram illustrating details of a motor control unit illustrated in FIG. 5;

FIG. 7A is a block diagram showing details of a motor driver unit in the case of a stepping motor, and FIG.
FIG. 3 is a block diagram showing details of a motor driver in the case of a servo motor.

FIG. 8 is a hydraulic circuit showing a shift control device according to one embodiment of the present invention.

[Explanation of symbols]

Reference Signs List 11 forward / reverse switching device 12 continuously variable transmission 14 primary pulley 16 secondary pulley 17 drive belt 18 primary oil chamber 21 secondary oil chamber 32 clutch cylinder 32a clutch oil chamber 34 forward clutch 45 brake cylinder 45a brake oil chamber 47 oil pump (high pressure 49) Oil pump (low pressure side pump) 50 Line pressure path 55 Clutch pressure regulating valve 81 Primary pressure path 82 Bidirectional pump 83 Motor 91 Control unit (pressure regulation control means, motor control means)

──────────────────────────────────────────────────続 き Continued on the front page (51) Int.Cl. ⁷ Identification symbol FI Theme coat ゛ (Reference) F16H 63:06 F16H 63:06

Claims (4)

[Claims] 1. The speed change of a belt-type continuously variable transmission having a primary pulley provided on a primary shaft driven by an engine and a secondary pulley provided on a secondary shaft and having a drive belt stretched between the primary pulley and the primary pulley. A control device, comprising: a primary cylinder that changes a pulley groove width of the primary pulley; a secondary cylinder that changes a pulley groove width of the secondary pulley; and one of the cylinders driven by the engine to regulate a discharge flow. And a positive displacement fluid machine driven in both forward and reverse directions by a motor to supply the discharge flow to the other cylinder. 2. The speed change of a belt-type continuously variable transmission having a primary pulley provided on a primary shaft driven by an engine and a secondary pulley provided on a secondary shaft and having a driving belt stretched between the primary pulley and the primary pulley. A control device, comprising: a primary cylinder that changes a pulley groove width of the primary pulley; a secondary cylinder that changes a pulley groove width of the secondary pulley; An oil pump to be supplied to the cylinder, and a bidirectional pump driven by a motor in both the forward and reverse directions to supply a discharge flow to the other cylinder using the regulated hydraulic pressure as an original pressure. Having a reversible rotating motor that rotates in both reverse directions. Shift control device of the transmission. 3. The speed change of a belt-type continuously variable transmission having a primary pulley provided on a primary shaft driven by an engine, and a secondary pulley provided on a secondary shaft and having a drive belt stretched between the primary pulley and the primary pulley. A control device, comprising: a primary cylinder that changes a pulley groove width of the primary pulley; a secondary cylinder that changes a pulley groove width of the secondary pulley; and a discharge flow driven by the engine to supply one of the cylinders. An oil pump; a pressure control means for controlling the hydraulic pressure of the hydraulic oil supplied to one of the cylinders in accordance with the traveling state of the vehicle; and a motor that is driven in both forward and reverse directions by a motor to return the regulated hydraulic pressure to the original pressure. A two-way pump that supplies a discharge flow to the other cylinder, A speed change control for a belt-type continuously variable transmission, comprising: a reversible rotary motor that rotates in both directions; and motor control means that controls the reversible motor in accordance with a running state such as an engine speed and a throttle opening. apparatus. 4. The shift control device for a belt-type continuously variable transmission according to claim 1, wherein the high-pressure side oil pump is connected to a high-pressure system oil passage connected to each of the cylinders. And a low-pressure-side pump is provided in a low-pressure oil passage provided separately from the high-pressure oil passage.