JPH0276966A - Oil pressure control device for belt type continuously variable speed change gear - Google Patents

Abstract

PURPOSE:To prevent a line pressure regulating value from being influenced by a flow rate of working oil flowing in a hydraulic servo by a method wherein from a predetermined relation, based on a control value of a flow rate to a flow rate control valve, a line pressure control value is corrected. CONSTITUTION:In a line pressure control valve correcting means, from a relation predetermined so that a line pressure responding to a line pressure control value by means of which a line pressure regulating valve 24 is driven is generated despite of a fluctuation in an inflow flow rate, the line pressure control value is corrected based on a control value of a flow rate to a flow rate control valve 30. Thereby, when by using the line pressure control value after correction, the line pressure regulating valve 24 is driven, a line pressure responding to the line pressure control value is generated despite of a fluctuation in working oil in the flow rate control valve 30. This constitution eliminates a fear of a belt nipping pressure between adjoining discs 7 and 8 on the input side and discs 9 and 10 on the output side being reduced according to a line pressure on the running condition of a vehicle on which a speed ratio is changed rapidly.

Description

DETAILED DESCRIPTION OF THE INVENTION Field of the Invention The present invention relates to a hydraulic control device for a belt-type continuously variable transmission used as a power transmission device for a vehicle.

2. Description of the Related Art The present applicant has previously proposed a power transmission system for a vehicle that uses a continuously variable transmission and can be operated with minimum fuel consumption over the entire operating range of the engine. Such power transmission devices generally include:
In a belt type continuously variable transmission equipped with an input side disk and an output side disk provided on the input shaft and the output shaft, and a belt wrapped around the input side disk and the output side disk, a predetermined level is set by a line pressure regulating valve. By adjusting the flow rate at which the hydraulic oil whose line pressure has been regulated to the line pressure flows into one of a pair of hydraulic servos that changes the effective diameters of the input side disk and the output side disk using a flow rate control valve, The speed ratio change speed of the two-belt continuously variable transmission is controlled. For example, the target value determined based on the actual throttle valve opening and vehicle speed from a predetermined relationship that allows the engine to operate along the minimum fuel efficiency curve and the actual engine rotational speed transformation or speed ratio. In speed ratio control that changes the speed ratio so that the speed ratios match, the speed ratio change speed is usually determined depending on the magnitude of the deviation between the target value and the actual value. Further, the line pressure is controlled by a line pressure regulating valve in relation to the actual speed ratio and transmission torque of the continuously variable transmission in order to maintain the belt tension at a necessary and sufficient value. .

Problems to be Solved by the Invention By the way, the line pressure regulating valve operates by leaking a part of the hydraulic oil that is force-fed from the oil pump to the line oil passage so that the line pressure corresponding to the line pressure control value is obtained. It is configured to regulate line pressure. However, in the line pressure regulating valve mentioned above, the relationship between the line pressure control value and the actual line pressure is established within the normal leakage amount range, but when the leakage amount decreases below that range, the line pressure control value will change. Since the pressure is regulated to a lower value than the specified value, when the hydraulic oil in the line oil passage is allowed to flow into one of the hydraulic servos through the flow control valve, the consumption of hydraulic oil in the line oil passage increases rapidly. As a result, the amount of hydraulic oil leaking from the line pressure regulating valve is significantly reduced, which may cause a phenomenon in which the line pressure decreases. Therefore, under vehicle running conditions where the speed ratio changes rapidly, there is a risk that the belt clamping force on the input side disk and the output side disk will decrease in relation to the line pressure.

FIG. 13 shows that in the conventional hydraulic control device, when the deviation in the speed ratio feedback control (for example, the difference ΔNi between the target rotational speed N i n " and the actual input shaft rotational speed N i n ) increases rapidly, In an attempt to eliminate the deviation, the flow rate Q of hydraulic oil flowing into the hydraulic servo through the flow control valve increases, and at the same time, the consumption amount of line pressure hydraulic oil increases, and at the same time, the output flow ff1Q in the line pressure regulating valve increases.

increases, and as a result of this increase, a large amount of hydraulic oil in the line oil passage is consumed, the amount of leakage at the line pressure regulating valve decreases, and the line pressure P2 is regulated lower than the value corresponding to the line pressure control value. The condition is shown. ΔPP indicates the difference between the value corresponding to the line pressure control value and the actual line pressure, that is, the line pressure reduction amount.

The present invention has been made against the background of the above circumstances,
The purpose is to provide a hydraulic control device for a belt-type continuously variable transmission in which a line pressure regulation value is not affected by the flow rate Q1 of hydraulic fluid flowing into a hydraulic servo through a flow control valve.

Means for Solving the Problems The gist of the present invention to achieve the object is to provide an input side disk and an output side disk provided on the input shaft and the output shaft, and an input side disk and an output side disk provided on the input side disk and the output side disk. In a belt-type continuously variable transmission equipped with a belt wrapped around the belt, hydraulic oil whose pressure is regulated to a predetermined line pressure by a line pressure regulating valve is used to change the effective diameters of the input side disk and the output side disk. A hydraulic control device for a belt type continuously variable transmission that controls the speed ratio change speed of the belt type continuously variable transmission by adjusting the inflow flow rate flowing into one of the hydraulic servos using a flow control valve. The flow rate control value to the flow rate control valve is determined based on a relationship determined in advance so that a line pressure corresponding to the line pressure control value for driving the line pressure regulating valve is obtained regardless of the inflow flow rate. It includes a line pressure control value correction means to correct the line pressure control value based on the line pressure control value.

Operation and Effect of the Invention By doing this, in the line pressure control value correction means,
Based on the flow rate control value to the flow rate control valve based on a relationship determined in advance so that a line pressure corresponding to the line pressure control value for driving the line pressure regulating valve can be obtained regardless of fluctuations in the inflow flow rate. Since the line pressure control value is corrected using the corrected line pressure control value, when the line pressure regulating valve is driven using the corrected line pressure control value, the line pressure control value corresponds to the line pressure control value regardless of fluctuations in the hydraulic fluid in the flow control valve. Line pressure is obtained. Therefore, under vehicle running conditions where the speed ratio changes rapidly, the possibility that the belt clamping force at the input side disk and the output side disk decreases in relation to the line pressure is eliminated.

EXAMPLE Hereinafter, an example of the present invention will be described in detail based on the drawings.

Figure 2 shows a constant fuel consumption rate line (solid line, unit: g/PS) in the two-dimensional coordinates of the engine rotation speed axis and the engine output torque axis.
-h) and isohorsepower lines (dashed lines, in PS) are shown. In the figure, the dashed line indicates the characteristic when the throttle valve is fully open, and is the operating limit of the engine. In addition, line A is a line connecting the points that give the minimum fuel efficiency for each output horsepower, and with conventional transmissions, the engine was operated along line B, making it impossible to obtain a sufficient fuel efficiency. It was. In the power transmission system of this embodiment, the required horsepower of the engine is set as a function of the operation amount of the accelerator pedal, that is, the throttle valve opening, and the engine rotation speed and engine output torque are positioned on line A at each required horsepower. The engine is operated. The engine rotational speed N1fi is controlled by changing the speed ratio of a continuously variable transmission (hereinafter referred to as CVT), and the engine output torque is controlled by changing the intake system throttle opening. The above function is determined on the assumption that the required horsepower increases as the amount of depression of the accelerator pedal increases.

In FIG. 3, an output shaft 2 of an internal combustion engine 1 is connected to an input shaft 5 of a CVT 4 via a clutch 3, and an output shaft 6 is disposed parallel to the input shaft 5. An input-side fixed disk 7 and an output-side fixed disk 9 are fixed to the input shaft 5 and output shaft 6, respectively, while the input-side movable disk 8 and the output-side movable disk 10 are spline-fitted or the like so that they cannot rotate relative to each other and are fixed to the axis. It is provided so that it can be moved in the direction. The pressure receiving area of the hydraulic servo for applying thrust to the input side movable disk 8 is made larger than the pressure receiving area of the hydraulic servo for applying thrust to the output side movable disk 10. 6, the position of the input side movable disk 8 with respect to the input side fixed disk 7 is opposite to the position of the output side movable disk 10 with respect to the output side fixed disk 9. fixed discs 7 and 9 and movable discs 8 and 10
The zero opposing surfaces are formed as conical surfaces that are spaced apart from each other toward the outer periphery, and a belt 11 for transmitting power is wound around a groove formed between them.

Therefore, as the hydraulic pressure acting on the hydraulic servo for applying thrust to the input side movable disk 8 and the hydraulic servo for applying thrust to the output side movable disk IO is changed, the speed ratio e ( =rotational speed of output shaft 6/rotational speed of input shaft 5) can be changed.

The rotational force output from the output shaft 6 of the CVT 4 is transmitted to the driving wheels via a forward/reverse switching device and a differential gear device (not shown).

The torque sensor 15 detects the torque of the input shaft 5, that is, the output torque T of the internal combustion engine 1, from changes in torsional stress or torsional angle in the input shaft 5.

The accelerator pedal sensor 16 detects the amount of depression of the accelerator pedal 18 by the driver's foot 17. The opening degree of the intake throttle of the internal combustion engine 1 is controlled by a throttle actuator 19 . The input side and output side rotation angle sensors 20.21 measure the rotation angle of the disk 7.10, respectively.
That is, the rotation speed is detected. The line pressure regulating valve 24 receives a relief flow M of hydraulic oil sent from the reservoir tank 26 via the oil passage 27 to the oil passage 28 by the oil pump 25.
By controlling Q, the line pressure P of the line oil passage 29
Adjust the pressure of ffi. Line pressure Pl is constantly supplied to the hydraulic servo of the output side movable disk 10 via the line oil passage 29. The flow rate control valve 30 controls the amount of hydraulic oil flowing into the hydraulic servo of the input side movable disk 8 and the amount of hydraulic oil flowing out from the hydraulic servo. For example, when maintaining the speed ratio e of the CVT 4 constant, a line oil passage 31 branching from the line oil passage 29, a drain oil passage 32, and an oil passage 33
By respectively blocking the input side movable disk 8, the axial position of the input side movable disk 8 is maintained constant. When increasing the speed ratio e of the CVT 4 (in the case of increasing speed change), the amount Q of hydraulic oil flowing from the line oil passage 31 to the oil passage 33 is increased to increase the thrust of the hydraulic servo of the input side movable disk 8. Conversely, when decreasing the speed ratio e of the CVT 4 (in the case of deceleration shifting), the amount of hydraulic oil flowing from the oil passage 33 to the drain oil passage 32 is increased and the hydraulic servo of the input side movable disk 8 is increased. reduce the thrust of Although the oil pressure in the oil passage 33 is lower than the line pressure PR, since the pressure receiving area of the hydraulic servo on the input end movable disk 8 is relatively large compared to the output side, the hydraulic servo on the input side movable disk 8 is It is possible to make the thrust larger than on the output side.

The electronic control unit 38 includes a D/A converter 40, an input interface 41 . A/D converter 42, CPU 43, RAM 4
4. Contains ROM45. The analog outputs of the torque sensor 15 and accelerator pedal sensor 16 are sent to the A/D converter 4.
2 and the pulses of the rotation angle sensor 20.21 are sent to the input interface 41. Throttle actuator 19, flow control valve 30, and line pressure regulating valve 24
The outputs from the D/A converter 40 to the amplifiers 49 .
50.51.

FIG. 4 shows an amplifier 49 for the throttle actuator 19.
shows the relationship between the input terminal and output current. FIG. 5 shows the relationship between the input current of the throttle actuator 19 and the intake system throttle opening. As a result, the throttle opening degree increases in proportion to the input voltage of the amplifier 49. FIG. 6 shows the relationship between the input voltage and output current of the amplifier 50 for the flow control valve 30, and FIG. 7 shows the relationship between the input current of the flow control valve 30 and the flow rate supplied to the input side hydraulic servo of the CVT 4. It shows. As is clear from this, if the differential pressure between the oil passages sandwiching the flow control valve 30 is constant, the speed ratio e changes at a speed proportional to the magnitude of the input voltage V in of the amplifier 50.

FIG. 8 shows the relationship between the input voltage and output current of the amplifier 51 for the line pressure regulating valve 24, and FIG. 9 shows the relationship between the input voltage and the output current of the line pressure regulating valve 24.
4 shows the relationship between input current and line pressure Pi. As is clear from this, the input voltage of the amplifier 51 is ■.

Line pressure Pi changes in proportion to the size of □. In addition,
Even if the input voltage to the line pressure regulating valve 24 is zero, the line pressure Pi is maintained at a predetermined value Pffil (Pffil≠1), so even if a disconnection or a failure occurs in the electronic control unit 38, the movable disk 8. Hydraulic oil is supplied to the IO hydraulic servo to ensure minimum torque transmission in the CVT4.

FIG. 10 is a block diagram of this embodiment.

Before explaining this diagram, the technical idea of this embodiment will be explained.

First, the input side movable disk 8 is passed through the flow control valve 30.
The flow rate Q of hydraulic oil flowing into the hydraulic servo that applies thrust to is normally expressed by the following equation (1).

Q+=Vi-礪會-・Se (1) However, VIn is the flow rate control value for the flow rate control valve 30 (
(voltage value) and ΔP are differential pressures generated on both sides of the flow control valve when hydraulic oil flows through the flow control valve.

Then, the following equation (2) is established from the above equation (1).

Q+ =K ・V+-' f (V.ut)"
-(2) However, K is a constant.

The specific formula for the above correction function f (V.ut) is CV
It is determined according to the mechanical properties and friction properties of T4.

Here, a constant flow rate control value ■8. ．． Even if the opening degree of the flow rate control valve 30 is constant in accordance with
1Q is increased and, conversely, the line pressure PR is decreased, the differential pressure ΔPi in equation (1) becomes smaller and the control flow ff
1Q, is decreased, so that the constant flow rate control value V
Since the control 4ffff flow ff1Q corresponding to i n cannot be obtained, conventionally, as shown in FIG. 12, a constant flow rate control value 1. ．． Even if the line pressure is P! The rate of change in speed ratio was not constant in relation to changes in . but,
In this embodiment, the flow rate control value V in is corrected using the deviation ratio (deviation ratio) m of the actual line pressure P2 with respect to a constant standard line pressure Pml determined in advance as a reference value, and after the correction By outputting the flow rate control value V in, the control flow rate Q1 corresponding to the initial flow rate control value V in
is now available.

That is, in this embodiment, the deviation ratio (deviation ratio) m of the actual line pressure PA with respect to the constant standard line pressure Pml is defined as the following equation (3), and the correction function f(■.ut
) is 1/m, the control flow rate Q is (4
) is expressed by the formula. Here, the pressure control value ■ is supplied to the line pressure regulating valve 24 and ultimately represents the line pressure approximately. Of ut, the above standard line pressure Pml
Assuming that the pressure control value supplied to the line pressure regulating valve 24 to generate is 1°, as shown in equation (3),
Approximately, the deviation ratio m is ■. It is expressed by ut/V-0. Therefore, as is clear from equation (4), for the control flow rate Q1, ■. U, will be involved in the form of a product.

m=Pf! , / P m E ζ■. ut/Vi. , ...(3) Q1=K
・■I, , /m ... (4) As is clear from the right side of equation (4) above, we can replace ■, , , /m with the new ■i
In other words, by using V r n corrected by dividing by m, the flow rate control value V, A control flow rate Q corresponding to A is obtained. Therefore, the flow rate control value V in for controlling the speed ratio change speed is determined according to the following equation (5). (5)
1/m on the right side of the equation is for correcting V i n .

Vmn=Kl (N'+n-Nzn)/m ・ ・
(5) On the other hand, the line pressure regulating valve 24 adjusts the line pressure P! by changing its relief flow rate Q! As shown in FIG. 11, the pressure control value V decreases as the relief flow ff1Q becomes smaller than the normal range. uL and the line pressure P which is set and maintained by pressure regulation. It is inevitable that the line pressure Pffi obtained will become lower due to a deviation in the relationship between the two. The output flow rate Q2 from the line pressure regulating valve 24 to the flow control valve 30 or the hydraulic servo is calculated using the following formula 1%. Since the flow rate is obtained by subtracting the relief flow rate Q of the line pressure regulating valve 24, for example, in order to rapidly change the speed ratio e to the speed increasing side, the hydraulic oil in the line oil passage 31 flows into the input side movable disk 8. If it is allowed to flow rapidly into the hydraulic servo,
Conventionally, by significantly reducing the relief flow rate Q, the line pressure was reduced to the pressure control value ■. could not be maintained at a value corresponding to ut. As shown in FIG. 13,
The deviation ΔN i n between the target rotation speed N', 7 in the speed ratio control and the actual rotation speed N i n of the input shaft 5 suddenly increases, and the output flow rate Q2 of the line pressure regulating valve 24 is rapidly increased. When this occurs, the relief flow IQ is significantly reduced, so that the line pressure PR shown by the solid line, as shown on the right side of the broken line, becomes the original line corresponding to the pressure control value V out, as shown by the dashed-dotted line. ΔP with respect to pressure Pl
It decreases by 2.

However, in this embodiment, the line pressure control value V out.

By calculating ΔP2 using equation (7), the pressure drop ΔP2 can be eliminated.

V, -t = to!・T-・Ni, /N0. t+Kl
・y in ... (7) That is, the above (7)
In the equation, the second term on the right side is the flow rate control value {circle around (2)} corresponding to the substantially actual controlled flow rate Q1 in the flow rate control valve 30. Since it is a term proportional to Correct by increasing the amount. For example, when a large amount of hydraulic oil flows into the hydraulic servo on the input side movable disk 8 through the flow rate control valve 30, the hydraulic oil is consumed based on the difference in diameter between the two hydraulic servos, and the line pressure regulating valve 2
Even if the relief flow ftQ of 4 deviates low from the set value (pressure value corresponding to the flow rate control value V in ) according to FIG. 11 due to a decrease, the deviation can be compensated for (7
) The flow rate control value V in is corrected according to the equation. Therefore, according to the corrected flow rate control value Vln, the flow rate control valve 3
By driving 0, a line pressure corresponding to the initial flow rate control value 1 can be obtained.

The block diagram in FIG. 10 mainly shows the electronic control device 3.
8 functions are shown. In the figure, in block 55, the accelerator pedal depression amount Xscc is converted into the required horsepower, that is, the target output horsepower PS'. Required horsepower PS
" is a function of X icc that increases with the increase of the pedal stroke i1 Based on the horsepower PS', the engineer's target rotational speed, that is, the target rotational speed N i of the input shaft 5
The relationship between the required horsepower PS' and the target rotational speed N i n " is shown by A in FIG. 2.

In other words, the engine rotation speed at which the required horsepower PS" can be obtained at the minimum fuel consumption rate is the target rotation speed. In block 57, the target rotation speed N, 1' and the actual rotation speed N in
Then, in the feedback gain of block 58, the flow rate control value v1 is set to 5) Calculated from Eq. This equation (5) is the deviation value m for correcting the flow rate control value V-.
Therefore, by using the corrected flow rate control value V in obtained by this equation (5), the actual rotational speed N,a can be quickly matched with the target rotational speed N,7゛. The control flow IQ is ensured regardless of the line pressure PR. In this embodiment, the block 58 is the flow rate control value V
It corresponds to a flow rate control value correction means for correcting rn.

On the other hand, in block 61, the actual output torque T8, the input shaft rotational speed N in , and the output shaft rotational speed N are calculated from the equation (7). ut, and the line pressure control value ■ based on the flow rate control value V in determined by equation (5). , is calculated. This equation (7) includes the line pressure control value ■. Since the correction term (second term on the right side) is included, the line pressure P2 is adjusted to the inflow flow ff1Q to the hydraulic servo by the corrected flow rate control value V in obtained by this equation (7). Line pressure control value regardless of ■. , is maintained stably at a value corresponding to . In this embodiment, the block 61 is the line pressure control value V
It corresponds to line pressure control value correction means for correcting 0□.

In addition, in FIG. 10, ■8. ．． and■. , is a flow rate control valve amplifier 50 and a line pressure regulating valve amplifier 5
1, the actual voltage output to the flow control valve '30 and the line pressure regulating valve 24 is multiplied by a constant. Therefore, the values obtained by multiplying the gains of the flow rate control valve amplifier 50 and the line pressure regulating valve amplifier 51 by the equation (5) and (force equation +, Kz, K) can be expressed as +
, Kz, and K3, ■i□ and ■ in equations (5) and (7). uL can be the actual output voltage to flow control valve 30 and line pressure regulating valve 24.

FIG. 1 is a flowchart illustrating the operation of the electronic control unit 38 having the functions shown in the block diagram of FIG. In the figure, in step 66, the depression 1x, ec of the accelerator pedal 18 is detected from the input from the accelerator pedal sensor 16, and the target output horsepower PS' corresponding to the depression amount X lee is calculated. In the following step 67, the target rotational speed N1fi' (=engine 1
(target rotational speed) is calculated, and in step 68, the deviation ratio m of the actual line pressure Pffi with respect to the constant standard line pressure Pm2 is calculated from equation (3). At this time, (2) is stored in advance to generate the constant standard line pressure Pml. , and ■ calculated in the previous cycle. , and are adopted. ■ Calculated in this previous cycle. 1. Although it has been corrected, it approximately corresponds to the actual line pressure.

In step 69, the flow rate control value V in is calculated from equation (5) based on the deviation ratio m. Then, in step 70, the output torque T1 of the engine l and the output shaft rotational speed N are determined. After uL is read, step s7
At t, the line pressure control value V is determined according to equation (7) above.

ut is calculated.

As described above, according to this embodiment, the line pressure P! From the relationship (5), which is determined in advance so that the flow rate corresponding to the flow rate control value V r n for driving the flow rate control valve 30 can be obtained regardless of fluctuations in the flow rate control valve 30, the standard line pressure Pml of the actual line pressure Pi is The deviation ratio m (=V, t/V, o
ut ), the flow rate control value V in is corrected, so when the flow rate control valve 30 is driven using the corrected flow rate control value V in , the line pressure P2 of the flow control valve 30 is The flow rate Q corresponding to the flow rate control value (1) can be obtained regardless of fluctuations in the flow rate. Therefore, it is possible to eliminate a decrease in followability or control responsiveness of the speed ratio control due to variations in the line pressure PQ depending on the running conditions of the vehicle.

Further, according to this embodiment, the actual output torque T, the human power shaft rotational speed N in , the output shaft rotational speed N6ut, and the flow rate control value V obtained from the equation (5) are obtained from the equation (7).
line pressure control value V based on i n. uL is calculated. Since this equation (7) includes a correction term (second term on the right side) for the line pressure control value V out, the corrected line pressure control value ■ obtained by this equation (7). .

As a result, the line pressure Pf is maintained stably regardless of the inflow flow rate Q1 to the hydraulic servo. For example, the flow control valve 30
As a large amount of hydraulic oil flows into the hydraulic servo on the input side movable disk 8 through According to FIG. 11, even if the line pressure control value deviates low from the set value (the pressure value corresponding to the line pressure control value V.ut), the line pressure control value ■ is set according to equation (7) so that the deviation can be compensated for. ut is corrected. Therefore, the line pressure control value after correction ■. , according to the line pressure regulating valve 2
4, the initial line pressure control value ■. ,
A line pressure corresponding to , can be obtained. Therefore, under vehicle running conditions where the speed ratio changes rapidly, the possibility that the belt clamping force at the input side disk and the output side disk decreases in relation to the line pressure is eliminated.

- The above is just one embodiment of the present invention, and the present invention can be modified in various ways using the knowledge of those skilled in the art without departing from its purpose.

[Brief explanation of the drawing]

FIG. 1 is a flowchart illustrating the operation of the electronic control device in the embodiment shown in FIG. FIG. 2 is a diagram showing equal horsepower lines and equal fuel consumption rate lines in two-dimensional coordinates of the rotational speed axis and the output torque axis of the engine. FIG. 3 is a block diagram showing the configuration of a vehicle power transmission device including an embodiment of the present invention. FIG. 4 is a diagram showing the input/output characteristics of the amplifier for the throttle actuator, and FIG. 5 is a diagram showing the relationship between the input of the throttle actuator and the throttle opening. FIG. 6 is a diagram showing the input/output characteristics of the amplifier for the flow control valve, and FIG. 7 is a diagram showing the relationship between the input of the flow control valve and the speed ratio of the CVT. FIG. 8 is a diagram showing the input/output characteristics of the line pressure regulating valve amplifier, and FIG. 9 is a diagram showing the input/output characteristics of the line pressure regulating valve. 10 and 11 are a functional block diagram and a flow chart showing the operation of the electronic control device in the embodiment of FIG. 3. 12th
The figure is a diagram showing the relationship between speed ratio change rate and line pressure in a conventional device. FIG. 13 is a time chart illustrating a conventional line pressure drop phenomenon. 4: CVT (belt type continuously variable transmission) 5: Input shaft 6: Output shaft 7: Input side fixed disk 8: Input side movable disk 9: Output side fixed disk IO = Output side movable disk 11: Belt 61 Niblock (flow rate Control value correction means) Applicant: Toyota Motor Corporation - Figure 4 Amplifier 49 (711 people Figure 5 OJ Luak call aviator 19 input 77 Electric - 7 Figure 6 Figure 7 Figure 8 O amplifier 51171λ power, 5 Figure 9 Line pressure localization valve 240 input V Figure 11 Line process Figure 12 Figure 13 Figure 0 Clear I=lt

Claims (1)

[Claims] A belt type continuously variable transmission comprising an input side disk and an output side disk provided on an input shaft and an output shaft, and a belt wound around the input side disk and the output side disk, A flow rate control valve controls the inflow flow rate of the hydraulic oil whose pressure has been regulated to a predetermined line pressure by the line pressure regulating valve into one of the pair of hydraulic servos that changes the effective diameter of the input side disk and the output side disk. A hydraulic control device for a belt type continuously variable transmission that controls the speed ratio change rate of the belt type continuously variable transmission by adjusting the speed ratio, the hydraulic control device for driving the line pressure regulating valve regardless of the inflow flow rate. line pressure control value correction means for correcting the line pressure control value based on the flow rate control value to the flow rate control valve from a relationship determined in advance so as to obtain a line pressure corresponding to the line pressure control value of the line pressure control value. A hydraulic control device for a belt type continuously variable transmission, which is characterized by: