JPH051762A - Duty solenoid control device for hydraulic actuation type transmission - Google Patents

Abstract

PURPOSE:To provide a duty solenoid control device which can improve the responsiveness of the oil pressure change against a change of duty ratio and the accuracy of the oil pressure control in a hydraulic actuation type transmission, in which oil pressure is controlled by a duty solenoid valve. CONSTITUTION:In this duty solenoid control device, a hydraulic actuation type non-stage speed change mechanism 10, a duty solenoid valve 52 for controlling the operating oil pressure of the non-stage speed change mechanism 10, and a control unit 33 for applying the duty ratio are provided. Provision of the frequency change means 33 (control unit 33) for increasing the driving frequency of the duty solenoid valve 52 when the control unit 33 changes the duty ratio is a characteristic of this control device. Desirably, the frequency change means 33 increases the driving frequency of the duty solenoid valve 52 only at the time of changing the duty ratio large.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a duty solenoid controller for a hydraulically actuated transmission.

[0002]

2. Description of the Related Art In a hydraulically actuated automatic transmission for an automobile, for example, a belt type continuously variable transmission, a driving pulley and a driven pulley whose pulley diameter can be freely changed, and a driving force from the driving pulley to the driven pulley are applied. A continuously variable transmission mechanism including a V-belt for transmission is provided. In such a continuously variable transmission, with respect to the drive pulley and the driven pulley,
A hydraulic piston mechanism for changing the pulley diameter is provided, and by controlling the hydraulic pressure applied to the hydraulic piston mechanism (piston chamber), the pulley diameters of both pulleys are adjusted to obtain a predetermined gear ratio. . The hydraulic pressure applied to the hydraulic piston mechanism is generally controlled by a duty solenoid valve that opens and closes according to a duty ratio applied from a control unit according to the engine speed, throttle opening, etc. (See JP-A-58-137652).

The shifting operation in such a continuously variable transmission mechanism is usually performed in the following procedure. That is, the engine speed by the control unit,
The duty ratio corresponding to the throttle opening etc. is calculated,
This duty ratio is applied to the duty solenoid valve, the duty solenoid valve is opened and closed according to this duty ratio, the amount of hydraulic oil released from the duty solenoid valve is adjusted, the hydraulic pressure applied to a predetermined hydraulic piston mechanism is adjusted, and continuously variable transmission is performed. The procedure is performed such that the gear ratio of the mechanism is adjusted.

[0004]

When it is necessary to change the hydraulic pressure applied to the hydraulic piston mechanism stepwise, for example, at the time of kickdown, in order to obtain good gear shifting characteristics, the engine speed and throttle It is necessary to quickly respond to changes in the hydraulic pressure supplied to the hydraulic piston mechanism to changes in the opening degree and the like. However, since the control unit calculates the duty ratio at a predetermined cycle, for example, every 25 ms, this causes a response delay or calculation delay (for example, 25 ms). In addition, the duty solenoid valve has a preset drive frequency.
Open / close pattern corresponding to the duty ratio in (Cycle)
(Pulse) is repeated, but if the duty ratio applied to the duty solenoid valve from the control unit is changed, the duty solenoid valve
At this point, after the cycle (opening / closing pattern) being executed ends, the opening / closing will be performed according to the changed duty ratio from the next cycle. Therefore, depending on the timing at which the duty ratio is applied from the control unit to the duty solenoid valve, there is a maximum response delay corresponding to the time required for the one cycle. For example, the drive frequency of the duty solenoid valve is 33H
In the case of z, a maximum response delay of about 30 ms will occur. For this reason, a delay in response to changes in hydraulic pressure with respect to changes in operating conditions may cause problems such as belt slip.

FIG. 5 shows a conventional belt type continuously variable transmission in which the duty ratio is changed stepwise from almost 100% to 0% at time t ₁ 'in order to perform rapid acceleration. An example of a change characteristic of time, line pressure, etc. In the example shown in FIG. 5, the time when the line pressure actually starts to rise is t ₂ ′, which is about 70 ms behind the time t ₁ ′. The time to increase the line pressure reaches 90% of the target on the booster has a t ₃ delayed further about 100ms than t ₂ '. Therefore, in this example, it is understood that there is a considerable delay in response, and there is a risk of causing problems such as belt slip due to insufficient belt clamping force. Needless to say, such a problem occurs not only in the case of a belt type continuously variable transmission but also in the case of an automatic transmission equipped with a transmission gear mechanism including a planetary gear system. Incidentally, if the line pressure is always kept high, the response of the hydraulic pressure change can be enhanced, but in this case, there is a problem that the power loss of the engine becomes large.

Therefore, for example, in an automatic transmission equipped with a speed change gear mechanism, when the load acting on the vehicle is large, the hydraulic system involved in the speed change gear mechanism is synchronized with the speed change timing of the speed change gear mechanism to a predetermined gear position. A duty solenoid control device has been proposed in which the drive frequency of the duty solenoid valve is increased to smooth the engagement of the clutch. However, in this conventional duty solenoid control device, the change in hydraulic pressure becomes smooth when the drive frequency is increased, and there is no idea to improve the responsiveness of the change in hydraulic pressure.
Further, if the drive frequency of the duty solenoid valve is increased, the number of times the duty solenoid valve is turned on and off is increased, which causes a problem that the durability thereof is deteriorated.

The present invention has been made to solve the above-mentioned conventional problems, and in a hydraulically actuated transmission in which the hydraulic pressure is controlled by a duty solenoid valve, the response of the hydraulic pressure change to the change of the duty ratio. It is an object of the present invention to provide a duty solenoid control device capable of increasing the hydraulic pressure control accuracy and increasing the accuracy of hydraulic control.

[0008]

To achieve the above object, a first aspect of the present invention is to provide a hydraulically actuated speed change mechanism, a duty solenoid for controlling the operating oil pressure of the speed change mechanism, and a duty ratio for the duty solenoid. In the hydraulically actuated transmission provided with the duty solenoid control means for applying, when the duty ratio applied to the duty solenoid by the duty solenoid control means is changed, the frequency changing means for increasing the drive frequency of the duty solenoid is provided. Provided is a duty solenoid control device in a hydraulically actuated transmission having the characteristics.

According to a second aspect of the present invention, in the duty solenoid control device for the hydraulically actuated transmission according to the first aspect, the frequency changing means increases the drive frequency of the duty solenoid only when the duty ratio change is large. There is provided a duty solenoid control device in a hydraulically actuated transmission.

[0010]

EXAMPLES Examples of the present invention will be specifically described below.
As shown in FIG. 1, a belt-type continuously variable transmission CT is provided for a four-cylinder engine 1 including first to fourth cylinders # 1 to # 4. Specifically,
The torque of the engine output shaft 2 is shifted at a predetermined gear ratio, and when the reverse range is selected, the rotation direction is reversed and transmitted to the drive wheels (not shown) via the gear train 3 and the differential device 4. It is supposed to do.

The belt type continuously variable transmission CT changes the torque of the engine output shaft 2 to change the torque of the turbine shaft 5.
To the primary shaft 7, the torque converter 6 that transmits the torque to the primary shaft 7, and the torque converter 6 that reverses the rotation of the turbine shaft 5 to transmit the torque to the primary shaft 7 when the reverse range is selected. And a belt-type continuously variable transmission mechanism 10 that transmits the signal to the transmission device 9. The secondary shaft 9
Are connected to the gear train 3.

The torque converter 6 is essentially a pump 1 connected to an engine output shaft 2 via a connecting member 11.
2 and a turbine 13 connected to the turbine shaft 5 and driven to rotate by hydraulic oil discharged from the pump 12.
And a stator 14 that rectifies the hydraulic oil that flows back from the turbine 13 to the pump 12 in a direction that promotes rotation of the pump 12.
The torque of the engine output shaft 2 is changed at a speed change ratio according to the difference in rotation speed between the pump 12 and the turbine 13. Here, the stator 14 is fixed to the transmission case 16 via a one-way clutch 15. A lockup clutch 17 that directly connects the engine output shaft 2 and the turbine shaft 5 is provided as needed.

The forward / reverse switching mechanism 8 is a planetary gear system. The forward / reverse switching mechanism 8 includes a sun gear 18 coaxially connected to the turbine shaft 2 and a ring gear 19 coaxially connected to the primal shaft 7. A plurality of pinion gears 20 that mesh with the sun gear 18 and the ring gear 19 and a carrier 21 that rotatably (rotatably) supports the pinion gears 20 are provided. The hydraulic clutch 2 is provided between the turbine shaft 5 and the carrier 21.
2 is provided, and a hydraulic brake 23 is provided between the carrier 21 and the transmission case 16 ′. Here, the hydraulic clutch 22 and the hydraulic brake 23 are respectively engaged when hydraulic pressure is supplied from a hydraulic mechanism FS described later.
(ON) and released when hydraulic pressure is released (OFF)
It is supposed to be done.

The forward / reverse switching mechanism 8 is in a neutral state when both the hydraulic clutch 22 and the hydraulic brake 23 are off, and the torque of the turbine shaft 5 is not transmitted to the primal shaft 7. Hydraulic clutch 22
When only one is turned on, the sun gear 18 and the carrier 21 are fixed to each other via the hydraulic clutch 22, so that the forward / reverse switching mechanism 8 loses freedom and each gear 18 to
20 and the carrier 21 rotate integrally. That is, the primal shaft 7 rotates in the same direction as the turbine shaft 5, and at this time, the vehicle can move forward. When only the hydraulic brake 23 is turned on, the carrier 21 is fixed to the transmission case 16 ′ via the hydraulic brake 23, so that the sun gear 18, the pinion gear 20, and the ring gear 19 form a fixed gear train that meshes in this order. Function. At this time, since the ring gear 19 rotates in the direction opposite to the sun gear 18, the primal shaft 7 rotates in the direction opposite to the turbine shaft 5, and the vehicle can move backward. At this time, the speed is reduced from the turbine shaft 5 to the primal shaft 7 at a gear ratio determined by the number of teeth of the sun gear 18 and the number of teeth of the ring gear 19. The hydraulic clutch 22 and the hydraulic brake 23
Of course, there is no case where both and are turned on.

The continuously variable transmission mechanism 10 substantially has a drive pulley 24 that rotates integrally with the primary shaft 7 and a driven pulley 25 that rotates integrally with the secondary shaft 9.
And a V-belt 26 that transmits the driving force of the drive pulley 24 to the driven pulley 25. Note that, hereinafter, for convenience, the engine side (right side in FIG. 1) when viewed in the axial direction of the primary shaft 7 is referred to as “right”, and the opposite direction is “left”.
Say. The drive pulley 24 is disposed on the left side of the first fixed conical plate 27 that is coaxially fixed to the primary shaft 7, and on the left side of the first fixed conical plate 27 so as to face the first fixed conical plate 27 so that the drive pulley 24 can move left and right with respect to the primary shaft 7. And the first movable conical plate 28. Then, the first movable conical plate 28
A first hydraulic cylinder 29 is provided to control the left and right positions of the. Here, when the hydraulic pressure applied to the first hydraulic cylinder 29 rises, the first movable conical plate 28 moves to the right,
The holding position of the V-belt 26 is displaced outward, and the drive pulley 2
When the effective pulley diameter of No. 4 increases and the hydraulic pressure decreases, the effective pulley diameter of the drive pulley 24 decreases. The driven pulley 25 is basically configured similarly to the drive pulley 24, and has a second fixed conical plate 30 coaxially fixed to the secondary shaft 9 and a right side of the second fixed conical plate 30 facing the second fixed conical plate 30. The second movable conical plate 31 is arranged.
Then, in order to control the left and right positions of the second movable conical plate 31, a second hydraulic cylinder 32 basically having the same structure as the first hydraulic cylinder 29 is provided.

In the continuously variable transmission mechanism 10, the effective pulley diameter of the drive pulley 24 is determined in accordance with the hydraulic pressure applied to the first hydraulic cylinder 29 from the hydraulic mechanism FS, while the second hydraulic cylinder 32 is driven by the driven pulley 24. The pulley 25 is V
The hydraulic pressure is applied so that the tension of the belt 26 can be constantly maintained at a predetermined value, that is, an effective pulley diameter that can transmit the driving force without causing belt slip can be obtained. Where the first
When the hydraulic pressure applied to the hydraulic cylinder 29 is increased,
Along with this, the effective pulley diameter of the drive pulley 24 increases. For this reason, the tension of the V-belt 26 tends to increase, but the hydraulic pressure is maintained in the second hydraulic cylinder 32 so that the tension of the V-belt 26 can be maintained at a predetermined value.
The effective pulley diameter of the driven pulley 25 is reduced so as not to increase the tension. In this way, the effective pulley diameter of the drive pulley 24 increases and the effective pulley diameter of the driven pulley 25 decreases, so that the gear ratio of the continuously variable transmission mechanism 10 continuously changes in the speed increasing direction. Incidentally, when the hydraulic pressure applied to the first hydraulic cylinder 29 is lowered, it goes without saying that the gear ratio of the continuously variable transmission mechanism 10 continuously changes in the deceleration direction contrary to the above case.

A hydraulic mechanism FS for supplying hydraulic pressure to the lockup clutch 17, the hydraulic clutch 22, the hydraulic brake 23, the first and second hydraulic cylinders 29, 32, etc. of the belt type continuously variable transmission CT is provided. The hydraulic pressure supply FS is controlled by the control unit 33 composed of a microcomputer. The control unit 33 is a comprehensive control device including the duty solenoid control means and the frequency changing means described in claims 1 and 2. The hydraulic mechanism FS will be described below.

As shown in FIG. 2, the hydraulic mechanism FS has a line pressure control section L for controlling the line pressure of the hydraulic mechanism FS.
And a gear ratio control unit M that controls the hydraulic pressure applied to the first hydraulic cylinder 29 of the continuously variable transmission mechanism 10. Further, although not shown, a lockup control unit that controls the hydraulic pressure applied to the lockup clutch 17 is provided. And the hydraulic clutch 22
And a forward / reverse switching control unit for controlling the hydraulic pressure applied to the hydraulic brake 23 (see FIG. 1).

The line pressure control unit L includes an oil pump 35.
A secondary valve 37 that adjusts the hydraulic pressure of the hydraulic oil discharged from the main hydraulic passage 36 to a predetermined hydraulic pressure (line pressure) corresponding to the pilot pressure is provided. This line pressure
(Operating oil) is supplied to each part of the hydraulic mechanism FS as a source pressure. The pilot pressure supplied to the secondary valve 37 is formed by the pressure reducing valve 38, the first duty solenoid valve 39 whose duty is controlled by the control unit 33, and the accumulator valve 40. Specifically, first, the line pressure (working oil) in the main oil passage 36 is introduced into the pressure reducing valve 38 via the line pressure supply passage 41, and the pressure reducing valve 3
At 8, the pressure is reduced to a predetermined hydraulic pressure and output to the pilot pressure passage 42. The hydraulic pressure in the pilot pressure passage 42 is
The first duty solenoid valve 3 which is opened and closed according to the duty ratio applied from the control unit 33.
Controlled by 9. The duty ratio is calculated by the CPU of the control unit 33 based on the engine speed, throttle opening, select range, etc.
It is calculated in a predetermined cycle (for example, once every 25 ms) and is output to the first duty solenoid valve 39.

The first duty solenoid valve 39 has a predetermined duty drive frequency or cycle number.
It operates at (for example, 33 Hz), and is opened in each cycle for a time corresponding to the duty ratio. For example, when the drive frequency is set to 33 Hz, each cycle is about 30 ms, but here the duty ratio is 40
%, 40% (12ms) before each 30ms cycle
Will be opened and 60% later (18 ms) will be closed and such a cycle will be repeated. Therefore, the higher the duty ratio, the more the hydraulic oil in the pilot pressure passage 42 is released, and the hydraulic pressure decreases.
The hydraulic pressure thus formed in the pilot pressure passage 42 becomes the pilot pressure. The pilot pressure passage 42
The hydraulic vibration or pulsation therein is absorbed by the accumulator valve 40. The pilot pressure thus formed is supplied to the first and second pilot oil chambers 43 and 44 of the secondary valve 37.

The secondary valve 37 has an axial direction
A spool 45 is provided which can be displaced (in the horizontal direction in FIG. 2). Hereinafter, for convenience, the left direction in FIG. 2 is simply referred to as “left”, and the opposite direction is referred to as “right”.
Although the spool 45 is divided into a large-diameter portion 45a and a small-diameter portion 45b, the spool 45 and the small-diameter portion 45b are always in contact with each other and are displaced substantially integrally. Therefore, these are not particularly distinguished below. The spool 45 is biased to the left by the biasing force of the spring 46 and the pilot pressure in the first and second pilot oil chambers 43 and 44. Therefore, when the pilot pressure decreases, the spool 45 is displaced to the right,
At this time, the input port 47 connected to the main oil passage 36 communicates with the drain port 48, and the oil pressure in the main oil passage 36 is increased.
(Line pressure) is released to the drain port 48, and the line pressure decreases. Conversely, when the pilot pressure rises, the spool 45 is displaced to the left, the communication area between the input port 47 and the drain port 48 becomes smaller, and the line pressure rises. In this way, the line pressure corresponding to the pilot pressure is formed. The line pressure is supplied to the second hydraulic cylinder 32 of the continuously variable transmission mechanism 10 via the second hydraulic cylinder oil passage 50 (see FIG. 1).

The gear ratio control section M includes a first hydraulic cylinder 2
A ratio valve 51 for directly controlling the hydraulic pressure applied to the valve 9 according to the control pressure;
A second duty solenoid valve 52 for controlling the control pressure applied to the engine, and a pitot valve 53 for switching between introducing the control pressure into the ratio valve 51 or the pitot pressure generated corresponding to the engine speed. ing.

A spool 54 is arranged in the ratio valve 51, and the spool 54 is biased rightward by a spring 55. A control oil chamber 56 is formed at the right end of the ratio valve 21, and a control pressure or pitot pressure is introduced into the control oil chamber 56 via a downstream control pressure passage 57. The ratio valve 51 has an input port 59 connected to a ratio valve passage 58 branched from the line pressure supply passage 41 and an output port connected to a hydraulic oil supply passage 60 for supplying hydraulic pressure to the first hydraulic cylinder 29. 61 and a drain port 62 are formed. If the hydraulic pressure in the control oil chamber 56 rises here, the spool 5
4 is displaced to the left, which increases the communication area between the input port 59 and the output port 61 and increases the hydraulic pressure applied to the first hydraulic cylinder 29. In this case, as described above, the effective pulley diameter of the drive pulley 24 increases, and the gear ratio of the continuously variable transmission mechanism 10 changes to the speed increasing side (see FIG. 1). On the contrary, when the oil pressure in the control oil chamber 56 decreases, the spool 54 is displaced to the right, which reduces the communication area between the input port 59 and the output port 61 or reduces the output port 61 and the drain port 62. The communication area between them increases, and the hydraulic pressure applied to the first hydraulic cylinder 29 decreases. In this case, the effective pulley diameter of the drive pulley 24 becomes smaller, and the gear ratio of the continuously variable transmission mechanism 10 changes to the deceleration side (see FIG. 1).

The second duty solenoid valve 52 is
Basically, the configuration is similar to that of the first duty solenoid valve 39, and the hydraulic pressure in the upstream control pressure passage 63 branched from the pilot pressure passage 42 is controlled according to the duty ratio applied from the control unit 33. , It is designed to form a control pressure.
Then, when the upstream control pressure passage 63 and the downstream control pressure passage 57 are connected by the pitot valve 53, the control pressure is supplied to the control oil chamber 56 of the ratio valve 51, and the hydraulic pressure applied to the first hydraulic cylinder 29 is increased. Is controlled by the second duty solenoid valve 52, that is, the control unit 33.
Will be controlled by. In addition, pitot valve 5
When 3 is shifted to the pitot pressure supply side, the downstream control pressure passage 57 is disconnected from the upstream control pressure passage 63 and is connected to the pitot pressure passage 64. At this time, the pitot pressure generated corresponding to the engine speed is supplied to the control oil chamber 56, and the gear ratio of the continuously variable transmission mechanism 10 (see FIG. 1) is controlled according to the pitot pressure.

By the way, as described above, a duty solenoid valve is generally accompanied by a response delay of a maximum of one cycle. Therefore, when the duty solenoid valve is used for hydraulic control of a continuously variable transmission, May cause a problem such as belt slip at the time of kickdown. Therefore, in this embodiment,
The control unit 33 controls the drive frequency of the second duty solenoid valve 52 to reduce the response delay without lowering the durability of the second duty solenoid valve 52.

The control method of the drive frequency control will be described below according to the flow chart shown in FIG. 3 and with reference to FIGS. Note that this control routine is executed at a predetermined cycle, for example, every 25 ms. Step #
In 1, the throttle opening, the engine speed, the hydraulic pressure P ₀ in the first hydraulic cylinder 29, etc. are read as control information. At step # 2, based on the throttle opening and the engine speed, as appropriate gear ratio in accordance with the operating state is obtained, the target hydraulic pressure P ₁ of the first hydraulic cylinder 29 is calculated.

At step # 3, it is determined whether or not the timer has timed out. This timer is reset (started) in step # 10 when the driving frequency is increased, and is set to time up after a predetermined time has elapsed. The predetermined time is set to be slightly longer than the time required for the hydraulic pressure to almost reach the target hydraulic pressure during the shift. If the timer times out in step # 3 (YES), step # 4
Then, it is compared and judged whether or not the absolute value | P ₀ −P ₁ | of the difference between the hydraulic pressure P ₀ and the target hydraulic pressure P ₁ exceeds the set value ΔP. In this embodiment, when | P ₀ −P ₁ | is large, it is necessary to suddenly change the gear ratio of the continuously variable transmission mechanism 10 due to kickdown (rapid acceleration) or the like, and the response of the hydraulic pressure change to the change in the operating state. If the performance is not improved, problems such as belt slip may occur. Therefore, the drive frequency of the second duty solenoid valve 52 is increased, the time required for one cycle thereof is shortened, and the response of the hydraulic pressure change is improved. There is. On the other hand, | P ₀ − even when shifting is required
When P ₁ | is small, the possibility of problems such as belt slip is small, so the drive frequency is not increased and the second
The durability of the duty solenoid valve 52 is improved.

If it is determined in step # 4 that | P ₀ −P ₁ | ≦ ΔP (NO), basically, in order to improve the durability of the second duty solenoid valve 52, step # 5 ~ The control routine for normal time of step # 6 is executed. Specifically, in step # 5, the drive frequency of the second duty solenoid valve 52 is the drive frequency f ₀ for normal time such that the drive frequency f ₀ for normal time does not cause deterioration of durability, for example, 33H.
z, and then at step # 6 this drive frequency f ₀
Is output to the second duty solenoid valve 52.
If it is determined in step # 3 that the timer has not timed out (NO), the driving frequency needs to be increased until the time is up, so steps # 4 to # 6 are skipped.

[0029] Thereafter, at step # 7, the valve characteristic of the second duty solenoid valve 52, taking into account the temperature compensation conditions, the duty ratio d ₁ required to obtain the target pressure P ₁ is calculated, step # In step 8, the duty ratio is changed to d ₁ (if the duty ratio d ₁ is the same as the previous time, it is not changed), and then in step # 9, this duty ratio d ₁ is applied to the second duty solenoid valve 52. After this, the process returns to step # 1.

On the other hand, in step # 4, │P ₀ -P ₁ │> Δ
If it is determined to be P (YES), the control routine for rapid gear shift of steps # 10 to # 12 is executed in order to enhance the responsiveness of the change in hydraulic pressure. Specifically, in step # 10, the timer is reset (started), and in step # 11, the driving frequency is set to a high driving frequency f ₁ that can sufficiently enhance the responsiveness, for example, 75 Hz, and then step # At 12, the drive frequency f ₁ is output to the second duty solenoid valve 52. After this, as in the case of the control routine for normal time described above, step # 7
~ Step # 9 is executed and the control unit 33
The duty ratio d ₁ is output from the second duty solenoid valve 52 to the second duty solenoid valve 52.

In this case, as described above, when the changed duty ratio d ₁ is applied to the second duty solenoid valve 52, the cycle based on the duty ratio before the change is applied.
If (one cycle) is being executed, the valve opening / closing operation according to the changed duty ratio d ₁ is not started until the end of this cycle. However, since the drive frequency is increased in advance, even if the changed duty ratio d ₁ is applied during the cycle execution, the time required to execute this cycle is shortened, and the duty ratio after the change is quickly changed.
The valve opening / closing operation according to d ₁ is started, and the responsiveness of the hydraulic pressure change to the change of the operating state is enhanced.

For example, as shown in FIG. 4, when the driving frequency is increased (G ₁ ), if the operating state is changed at time t ₁ , the second duty solenoid valve 5 is changed at time t ₂ .
The drive frequency of No. 2 is switched from f ₀ (low frequency) to f ₁ (high frequency), and at time t ₃ , the second duty solenoid valve 5
The changed duty ratio d ₁ is applied to 2, and the valve opening / closing operation according to the changed duty ratio d ₁ is started at time t ₄ . Therefore, the response delay in this case is D ₁ . On the other hand, when the drive frequency is not increased as in the normal time control routine (G ₂ ), the time when the changed duty ratio d ₁ is applied to the second duty solenoid valve 52 is G _Although it is t ₃ as in the case of ₁ , the second duty solenoid valve 52 actually starts the valve opening / closing operation at time t ₅ . The response delay in this case is D ₂ . Therefore, in the case of the control routine for a sudden shift, it can be seen that the responsiveness is significantly improved.

After that, the process returns to step # 1. When the driving frequency is increased in this way, steps # 3 to # are performed until the timer times out in step # 3.
Since the process skips to 7, the state in which the drive frequency is increased continues for a predetermined time. Thus, in this embodiment,
Only when the required hydraulic pressure change due to changes in operating conditions is large,
Since the drive frequency of the second duty solenoid valve 52 is increased, the responsiveness of the hydraulic pressure change to the change of the operating state can be improved without lowering the durability.

Although the present invention is applied to the belt type continuously variable transmission in this embodiment, the present invention is not limited to this, and for example, an automatic transmission provided with a speed change gear mechanism. Of course, it can also be applied to. That is, in an automatic transmission equipped with a speed change gear mechanism, even when a clutch or the like is engaged,
This is because it is necessary to perform hydraulic control according to the input torque from the engine in order to obtain a predetermined clutch static capacity, and the present invention can be effectively applied to such control.

[0035]

In general, in the duty solenoid, even if the applied duty ratio is changed, the duty ratio after the change is applied until the cycle (one cycle) at the time when the duty ratio is applied is completed. According to the first aspect of the invention, the drive frequency of the duty solenoid valve is increased when the duty ratio to be applied to the duty solenoid is changed. The time required for one cycle is shortened and the response delay is reduced. Further, since the responsiveness is enhanced in this way, the line pressure can be set relatively low, and the power loss is reduced.

According to the second invention, basically the same action and effect as the first invention can be obtained. Furthermore, since the drive frequency is increased only when the amount of change in the duty ratio is large, the drive frequency cannot be increased unnecessarily when the responsiveness does not need to be increased. The durability of the valve does not deteriorate. That is, it is possible to achieve both improvement in responsiveness and improvement in durability.

[Brief description of drawings]

FIG. 1 is a system configuration diagram of a belt type continuously variable transmission according to the present invention.

FIG. 2 is a system configuration diagram of a hydraulic mechanism of the belt type continuously variable transmission shown in FIG.

FIG. 3 is a flowchart showing a control method of drive frequency control.

FIG. 4 is a diagram showing operating characteristics of a duty solenoid valve that controls the hydraulic pressure applied to the first hydraulic cylinder of the continuously variable transmission mechanism.

FIG. 5 shows a conventional duty solenoid valve,
It is a figure which shows the characteristic of the hydraulic pressure change with respect to the change of a duty ratio.

[Explanation of symbols]

CT: Belt type continuously variable transmission FS ... hydraulic mechanism 10 ... continuously variable transmission 29 ... 1st hydraulic cylinder 33 ... Control unit 39 ... 1st duty solenoid valve 52 ... Second duty solenoid valve

Claims (2)

[Claims] 1. A hydraulically actuated transmission provided with a hydraulically actuated speed change mechanism, a duty solenoid for controlling the operating oil pressure of the speed change mechanism, and a duty solenoid control means for applying a duty ratio to the duty solenoid. , When the duty solenoid control means changes the duty ratio applied to the duty solenoid,
A duty solenoid control device for a hydraulically actuated transmission, characterized by comprising frequency changing means for increasing the drive frequency of the duty solenoid. 2. The duty solenoid control device for a hydraulically actuated transmission according to claim 1, wherein the frequency changing means increases the drive frequency of the duty solenoid only when the duty ratio change is large. And a duty solenoid control device in a hydraulically actuated transmission.