JPS6338042A - Hydraulic pressure control device for belt type continuously variable transmission for vehicle - Google Patents

Abstract

PURPOSE:To reduce the driving loss of a pump leading to the power loss of an engine by making the steady-state deviation between an actual speed ratio and a target speed ratio a controlled variable and feedback controlling a pressure regulating valve for regulating a line pressure so that the steady-state deviation agree with a target deviation value. CONSTITUTION:A speed change control valve 44 varies the effective diameters of speed change pulleys 20, 22 to control a speed ratio, by making feeding/discharging control of a first line hydraulic pressure which is discharged from an oil pump 42 and regulated by a first pressure regulating valve 48, to and from the hydraulic cylinders of the primary side and secondary side variable pulleys 20, 22. And, the first pressure regulating valve 48 is feedback controlled to regulate the first line hydraulic pressure in order to make the steady-state deviation between an actual speed ratio and a target speed ratio agree with a defined target deviation value. Also, the line hydraulic pressure at the time when the steady-state deviation nearly agrees with the target deviation value is stored being related to a vehicle operating condition, to carry out the control of the first pressure regulating valve 48 based on the stored value when a similar operating condition takes place afterward.

Description

DETAILED DESCRIPTION OF THE INVENTION Technical Field The present invention relates to a hydraulic control device for a belt-type continuously variable transmission for vehicles, and in particular to a hydraulic control device for controlling a first line hydraulic pressure on a high pressure side based on a steady deviation of a speed ratio. This relates to a control device.

Prior Art A pair of primary variable pulleys and a secondary variable pulley provided on the primary rotating shaft and the secondary rotating shaft, respectively, and a power transmission belt that is wound around the pair of variable pulleys to transmit power; 2. Description of the Related Art A belt-type continuously variable transmission for a vehicle is known that includes a pair of primary and secondary hydraulic cylinders that respectively change the effective diameters of a pair of variable pulleys.

A hydraulic control device for such a vehicle belt-type continuously variable transmission includes (a) a first pressure regulating valve that regulates the pressure of hydraulic fluid supplied from a hydraulic source to a first line hydraulic pressure; (b) the first line hydraulic pressure;
By supplying hydraulic oil whose pressure has been regulated to the line hydraulic pressure to one of the downstream side hydraulic cylinder and the secondary side hydraulic cylinder, and at the same time causing the hydraulic oil in the other side to flow out, the downstream side variable pulley and the secondary side a speed change control valve that adjusts the speed ratio of the continuously variable transmission by changing the effective diameter of the variable pulley;
c) A second pressure regulating valve that regulates the pressure of the hydraulic oil flowing out from the other of the upstream side hydraulic cylinder and the secondary side hydraulic cylinder through the speed change control valve, and sets the pressure of the hydraulic oil to a second line hydraulic pressure lower than the first line hydraulic pressure. A system has been considered in which the speed change control valve is controlled so that the actual speed ratio matches the target speed ratio determined according to the driving state of the vehicle. For example, the patent application filed earlier by the applicant in 1986-3
The device described in No. 7576 is one example.

In such a hydraulic control device, the first line hydraulic pressure and the second line hydraulic pressure are regulated by the first pressure regulating valve and the second pressure regulating valve, respectively, so that the first line hydraulic pressure has a speed ratio change speed, that is, a speed change response. The first pressure regulating valve is controlled so that the second line oil pressure is a necessary and sufficient value that does not cause power loss, while the second line hydraulic pressure is controlled so that the second line oil pressure is a necessary and sufficient value that does not cause slippage of the transmission belt. By controlling the two pressure regulating valves, sufficient transient response characteristics for speed ratio control can be obtained while maintaining vehicle power loss as low as possible.

Problems to be Solved by the Invention By the way, in such a conventional hydraulic control device, the first
When adjusting the line oil pressure, first find the thrust of the variable driven pulley that provides the belt squeezing force necessary for torque transmission, and then calculate the thrust value and the predetermined thrust according to the target speed ratio and engine output torque. The thrust of the drive-side variable pulley is calculated based on the ratio, that is, the ratio of the thrust of the driven-side variable pulley to the thrust of the drive-side variable pulley. Next, the required oil pressure for the drive side hydraulic cylinder is calculated from the thrust force calculation value of the drive side variable pulley, and the first line oil pressure is determined by adding the surplus oil pressure to this oil pressure calculation value. The first pressure regulating valve is controlled so as to obtain the desired pressure.

Here, the above-mentioned surplus oil pressure is necessary for reducing the steady deviation of the speed ratio. That is, the output oil pressure characteristics of the speed change control valve are as shown in FIG. 4, for example, and a speed ratio is achieved when the oil pressure in both hydraulic cylinders is at the oil pressure indicated by the circle. If so, there is a difference Δ■ between that speed ratio and the target speed ratio. A steady deviation of a magnitude corresponding to the above occurs, but this steady deviation becomes smaller when the first line oil pressure PJ is increased because the slope of the hydraulic characteristics becomes steeper, and when the first line oil pressure Pa1 is decreased, the steady deviation becomes smaller. This increases because the slope of the hydraulic characteristics becomes gentler. However, as the first line oil pressure pH increases, the driving loss of the pump also increases accordingly, so the above-mentioned margin oil pressure is determined at a balance point between the driving loss and the steady-state deviation, which are contradictory to each other.

On the other hand, due to individual differences in each component that constitutes a belt-type continuously variable transmission, the predetermined thrust ratio characteristics and the actual characteristics of the continuously variable transmission are not necessarily consistent, and the actual The thrust ratio characteristics of a multi-stage transmission change over time due to changes in sliding resistance, deterioration of lubricating oil, etc., but it is extremely difficult to set the thrust ratio characteristics in anticipation of such changes over time. be. Further, due to accuracy problems of the pressure regulating system including the 111th pressure valve and the second pressure regulating valve, the first line hydraulic pressure and the second line hydraulic pressure are not necessarily regulated as calculated.

Therefore, if you try to control the steady-state deviation of the speed ratio so that it is always smaller than a certain value, the above-mentioned error in the hydraulic pressure calculation value of the drive-side hydraulic cylinder, the first line hydraulic pressure, and the second line hydraulic pressure
It was necessary to set the above-mentioned margin oil pressure to a large value in consideration of pressure adjustment errors in the line oil pressure. For this reason, an unnecessarily high first line oil pressure is prepared in the operating range with few errors or when no changes occur over time, which causes drive loss of the pump and even power loss of the engine, resulting in vehicle fuel efficiency. There was an inconvenience in that it would be damaged.

Means for Solving the Problems The present invention has been made against the background of the above circumstances.
The gist is that (a) the first yA pressure valve;
(b) a speed change control valve; and (C) a second XV4 pressure valve, and controls the speed change control valve so that the actual speed ratio is equal to the target speed ratio determined according to the driving state of the vehicle. In the hydraulic control device of the type, the first line hydraulic pressure is regulated so that the steady deviation between the actual speed ratio and the target speed ratio is equal to a predetermined target deviation value.
While performing feedback control on the 1-pressure valve, the 1st line oil pressure at the time when its steady deviation is approximately equal to its target deviation value is stored in relation to the operating state of the vehicle, and when the same operating state subsequently occurs, the first i1 based on the stored value without performing the feedback control;
The present invention is provided with a control means for controlling the one-pressure valve.

Effects of the invention In other words, the present invention provides the first pressure regulating valve, the speed change control valve, and the second! In a hydraulic control device having a Ji pressure valve, we focused on the fact that the larger the first line oil pressure is, the smaller the steady deviation between the actual speed ratio and the target speed ratio becomes, and the steady deviation is used as a control variable to match the target deviation value. 1st U to lose! 4
The first line oil pressure is regulated by feedback controlling the pressure valve. By doing this, the steady deviation will be matched with the target deviation value, and
The first line oil pressure is controlled to the minimum oil pressure necessary to achieve the speed ratio that includes the target deviation value, so it is controlled to the oil pressure that is added to the excess oil pressure in anticipation of various errors, etc., as in the past. Compared to the case where the pump is operated, the drive loss of the pump and further the power loss of the engine are reduced, and the fuel efficiency of the vehicle is improved.

Here, the above-mentioned target deviation value is determined by, for example, a constant speed ratio steady deviation that is optimal for achieving a predetermined purpose through experiments or simulations in the overall driving state of the vehicle such as when driving in 10 modes. , even if the constant value is set as the target deviation value, or the minimum value of the steady-state deviation required in various operating conditions is set as the target deviation value, or the speed ratio of the continuously variable transmission, the output of the engine, etc. torque.

A data map provides the optimal target deviation value according to the vehicle driving condition determined by engine speed, etc.

Various aspects can be adopted, such as finding it using an arithmetic expression or the like.

In addition, in this way, the actual first line oil pressure is always controlled to the minimum necessary oil pressure regardless of pressure adjustment errors in the first line oil pressure and the second line oil pressure, so the 1U4 pressure valve and the 2nd There is no problem even if the pressure regulation accuracy of the pressure regulation system including the A pressure valve is low, and the steady deviation of the speed ratio also depends on the characteristics of the speed change control valve, but based on the steady deviation of the actual speed ratio, Since the line oil pressure is regulated, it will not be affected even if there are variations in the characteristics of the speed change control valve.Therefore, it is not necessary to use highly accurate pressure regulating systems and speed change control valves. The manufacturing cost can be reduced.

On the other hand, when the speed ratio steady-state deviation is made to substantially match the target deviation value in this way, the first line oil pressure at that time is stored in relation to the operating state of the vehicle, and the same operating state is subsequently achieved. Sometimes, the first pressure regulating valve is controlled based on the stored value. Therefore, after the optimal first line oil pressure is achieved under various operating conditions,
The first line oil pressure is immediately adjusted to the minimum oil pressure necessary to achieve the speed ratio that includes the target deviation value.
Weighted and more efficient reduction in fuel consumption is achieved compared to when pressure is regulated by constant feedback control.

In particular, if feedback control is performed during gear shifting, the first line oil pressure becomes extremely large, so it is necessary to limit or stop the feedback function during gear shifting. There is an advantage that the first line oil pressure can be regulated to the minimum necessary oil pressure.

Note that when storing the first line oil pressure, the first
It is desirable to add the oil pressure increase due to centrifugal force to the line oil pressure and store it in a two-dimensional map of speed ratio and engine output torque. This is because even if the speed ratio and engine output torque are the same, if the rotational speed of the variable pulley is different, the thrust due to centrifugal force will be different, and the optimal first line oil pressure will also change, so the value of the first line oil pressure is This is because, if the data is stored as is, a three-dimensional map including the rotational speed of the variable pulley must be used, which is undesirable as it requires a large capacity memory.

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

In FIG. 1, the output of an engine 10 installed in a vehicle is transmitted to a primary rotating shaft 16 of a belt type continuously variable transmission 14 via a clutch 12.

The belt type continuously variable transmission 14 includes a primary side rotating shaft 16 and a secondary rotating shaft 18, and a primary side variable transmission wheel with a variable effective diameter attached to the downstream rotating shaft 16 and the secondary rotating shaft 18. The pulley 20 and the secondary variable pulley 22, and the
A transmission belt 24 that is wound around the next variable pulley 20 and the secondary variable pulley 22 to transmit power, and a primary hydraulic cylinder 26 that changes the effective diameter of the next variable pulley 20 and the secondary variable pulley 22. and a secondary side hydraulic cylinder 28. These next hydraulic cylinders 26
The secondary hydraulic cylinders 28 are formed to have the same pressure receiving area, and the outer diameters of the secondary variable pulley 20 and the secondary variable pulley 22 are the same, so that the belt-type continuously variable transmission 14 is small. The above-mentioned secondary side variable pulley 20 and secondary side variable boolean IJ 22 are as follows:
- Fixed rotating bodies 31 and 32 fixed to the next rotating shaft 16 and the secondary rotating wheel 18, respectively; The fixed rotating body 31 may be provided
and 32, and movable rotating bodies 34 and 36 forming grooves between them.

The output from the secondary rotating wheel 18 of the belt-type continuously variable transmission 14 is transmitted to the drive wheels of the vehicle via an auxiliary transmission, a differential gear device, etc. (not shown).

A hydraulic control circuit for operating the vehicle power transmission device configured as described above is configured as described below. That is, the oil tank 38 passes through a return path (not shown).
The hydraulic oil returned to the strainer 40 and the suction oil passage 41
The oil is sucked into the oil pump 42 via the transmission control valve 44.
is fed under pressure to the first line oil passage 50 connected to the input port 46 and the first pressure regulating valve 4B. This oil pump 42
constitutes the hydraulic power source of this embodiment, and is driven by the engine 10 via a drive shaft (not shown).

The first pressure regulating valve 48 controls the hydraulic pressure in the first line oil passage 50 by causing a part of the hydraulic oil in the first line oil passage 50 to flow out to the second line oil passage 52 in accordance with a first drive signal VDI, which will be described later. (1st line oil pressure). The second line oil passage 52 is connected to a first discharge port 54 and a second discharge port 56 of the speed change control valve 44, and a second pressure regulating valve 58, respectively. The second pressure regulating valve 58 receives a second drive signal V, which will be described later.
By causing a part of the hydraulic oil in the second line oil passage 52 to flow out to the drain oil passage 60 according to D2, the oil pressure in the second line oil passage 52 (second line oil pressure) is made lower than the first line oil pressure. is also controlled to a relatively low value. The first pressure regulating valve 48 and the second pressure regulating valve 58 are constructed from so-called electromagnetic proportional relief valves.

The speed change control valve 44 is a so-called proportional control solenoid valve, and the input port 46. The first discharge port 54 and the second
Discharge boat 56. It communicates with a pair of first output capotros 2 and second output capotros 4, which are connected to the downstream side hydraulic cylinder 26 and the secondary side hydraulic cylinder 28 via connection oil passages 29 and 30, respectively. A cylinder bore 66 is formed in the valve body 65, one spool valve element 68 is slidably fitted into the cylinder bore 66, and the spool valve element 68 is biased toward a neutral position from both ends thereof. A pair of first springs 70 and a second spring 72 are provided at both ends of the spool valve 68 to hold the spool valve 68 in a neutral position. It includes a first electromagnetic solenoid 74 and a second electromagnetic solenoid 76 that are continuously moved against the biasing force of the first spring 70. The above spool valve 68
Four lands 78, 80, 82, 84 are formed sequentially from one end, and a pair of lands 80 and 82 located in the middle are arranged so that the axis of the spool valve 68 is aligned when the spool valve 68 is in the neutral position. It is formed at the same position as the first output capotro 2 and the second output capotro 4 in the direction. Further, the inner peripheral surface of the cylinder bore 66 is located at a position facing the pair of lands 80 and 82 when the spool valve element 68 is in the neutral position, that is, the first output capotro 2 and the second output capotro 4. is cylinder bore 6
A pair of first annular grooves 86 with a width slightly larger than the lands 80 and 82 are provided at positions opening into the inner circumferential surface of the lands 80 and 82.
and a second annular groove 88 are formed. The first annular groove 86 and the second annular groove 88 form a constriction whose flow cross-sectional area changes continuously in order to control the flow of hydraulic oil between the lands 80 and 82.

As a result, when the spool valve 68 is in the neutral position, the first output capotro 2 and the second output capotro 4
are evenly communicated with the input port 46 and the discharge port 54, 56 with a small circulation area, and an amount of hydraulic oil sufficient to replenish leakage is supplied to the primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28. Also, a small amount of hydraulic oil is drained from the discharge boat 54,56.

However, as the spool valve element 68 is moved from the neutral position in its uniaxial direction, for example, in the direction approaching the second electromagnetic solenoid 76 (i.e., rightward in the figure), the first output capotro 2 and the first discharge While the flow cross-sectional area with the port 54 is continuously increased, the flow cross-sectional area between the second output port 4 and the input port 46 is continuously increased. The working oil pressure output to the side hydraulic cylinder 26 is lower than the working oil pressure output from the second output capotro 4 to the secondary side hydraulic cylinder 28. For this reason, the balance between the thrust forces of the primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28 in the belt-type continuously variable transmission 14 is disrupted, and while hydraulic fluid flows into the secondary hydraulic cylinder 28, The hydraulic oil in the cylinder 26 flows out, and the speed ratio e of the belt-type continuously variable transmission 14 (rotational speed N.ut of the secondary rotating wheel 18/-rotational speed N1n of the secondary rotating shaft 16) decreases.

On the other hand, as the spool valve element 68 is moved from the neutral position toward the first electromagnetic solenoid 74, that is, to the left in the figure, the flow cross-sectional area between the first output capotro 2 and the input port 46 decreases. is continuously increased, while the flow cross-sectional area between the second output capotro 4 and the second discharge port 56 is continuously increased. The output hydraulic pressure is the second
The hydraulic pressure is higher than the working pressure output from the output capotro 4 to the secondary hydraulic cylinder 28. For this reason, the balance between the thrust forces of the primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28 in the belt-type continuously variable transmission 14 is disrupted, so that while hydraulic oil flows into the secondary hydraulic cylinder 26, the secondary hydraulic The hydraulic oil in the cylinder 28 flows out, and the speed ratio e of the belt type continuously variable transmission 14 increases. In this way, the speed change control valve 44 functions as a switching valve that supplies high-pressure hydraulic oil to one of the hydraulic cylinders 26 and 28 and low-pressure hydraulic oil to the other, and continuously adjusts the flow rate of the hydraulic oil. It also has a flow control valve function.

In addition, the second line oil passage 52 and the primary side hydraulic cylinder 2
6 (connection oil passage 29), a throttle oil passage 112 equipped with a throttle 110 for restricting the flow of hydraulic oil is connected to the second line oil passage 52 and the secondary side hydraulic cylinder 28. (Connection oil passage 30) is connected with a throttle oil passage 116 provided with a restriction 114 that restricts the flow of hydraulic oil. As a result, the hydraulic cylinder 26 on the high pressure side (drive side)
28 to the second line oil passage 52, the output oil pressure of the speed change control valve 44 (-the oil pressure p,, l of the downstream hydraulic cylinder 26 and the secondary oil pressure The oil pressure P of the cylinder 28 (□) is the spool valve 68
is brought close to the second line hydraulic pressure Plt at the neutral position. Therefore, the hydraulic cylinder 2 on the low pressure side (driven side)
The oil pressure P4 or P out in the line 6 or 28 is made to substantially match the second line oil pressure Pβ2.

On the other hand, the belt-type continuously variable transmission 14 has a negative rotation shaft 16.
A first rotation sensor 90 for detecting the rotation speed N, 7 of
, and the rotational speed N of the secondary rotating wheel 18. A second rotation sensor 92 is provided for detecting the rotation speed N, and the first rotation sensor 90 and the second rotation sensor 92 output a rotation signal SRI representing the rotation speed N, .
. A rotation signal SR2 representing ut is output to the controller 94. Further, the engine 10 is provided with a throttle sensor 96 for detecting the throttle valve opening θ as a quantity representing the required output of the vehicle, and an engine rotation sensor 98 for detecting the engine rotation speed N. , the throttle sensor 96 and the engine rotation sensor 98
From is a throttle signal S representing the throttle valve opening θth.
A rotation signal SE representing θ and engine rotational speed N1 is output to the controller 94.

The controller 94 includes the CPU 102. ROM104
．． This is a so-called microcomputer including RAMI O6 and the like, and constitutes the control means of this embodiment. Above CPU
102 uses the memory function of the RAM 106 to store R in advance.
The input signal is processed according to the program stored in the OMI 04, and the first drive signal VDI and the second drive signal are sent to the first pressure regulating valve 48 and the second pressure regulating valve 58 to control the first line oil pressure and the second line oil pressure. At the same time as respectively supplying the signal VD2, speed ratio signals RAI and RA2 are supplied to the first electromagnetic solenoid 74 and the second electromagnetic solenoid 76 for driving them in order to control the speed ratio e.

The operation of this embodiment will be explained below with reference to the flowcharts of FIGS. 2 and 3.

First, by executing step S1, the rotational speed N, 1 . Rotational speed N of the secondary rotating wheel 18. uL + Throttle valve opening θ, h.

The engine rotation speed N is the rotation signal SRI and SR2,
Based on the throttle signal Sθ1 rotary sign SE
Loaded on 06. Next, in step S2, R
The actual speed ratio e is calculated from the rotational speed N in and N.uL according to the following equation (11) stored in the OM 104. In addition, in step S3, the throttle Based on the valve opening degree θ, a target rotational speed N in ” of the primary rotation shaft 16 is determined. This is, for example, as shown in FIG.
This is determined in advance so that the engine 10 operates exclusively on the minimum fuel consumption rate curve shown in FIG. Then, in step S4, the target rotational speed N i n and the actual rotational speed N are determined according to the following equation (3). Calculate the target speed ratio all from □.

e = NouL/N = n...(1)N
=n"=f(θ)...(2)e"=
N out/ N t', wealth 40. (
3) In the following step S5, the speed ratio control value ■ is determined according to the following equation (4) stored in the ROM 104 in advance. Calculate.

In step S26, which will be described later, this speed ratio control value ■
. If is positive, the spool valve 68 is moved to the left and the rotational speed N of the secondary rotating wheel 18 is increased. The speed ratio signal RA2 is outputted so that ut increases, and when it is negative, the spool valve element 68 is moved to the right and the rotational speed N in of the primary rotating shaft 16 increases. A speed ratio signal RAI is output. Also, the speed ratio control value ■.

The magnitude corresponds to the magnitude of the speed ratio signal RAI or the speed ratio signal RA2, that is, the amount of movement of the spool valve 68. Therefore, as is clear from the following equation (4), the speed ratio control value ■. is determined so that the actual speed ratio e matches the target speed ratio ex. Note that K in equation (4) is a control constant.

V, -K(e”-e)/e ・-・<4) In the following step S6, the following equation (5
), the actual output torque T of the engine 10 is determined based on the throttle valve opening θ and the engine rotational speed N.
, is determined.

T,=f(θ, No) (5) Then, in step S7, it is determined whether the actual output torque T0 of the engine 10 is positive or not, that is, the positive torque state where power is output from the engine 10. It is determined whether the engine is braking or not. The reason why such a judgment is necessary is because the power transmission direction is different between the positive torque state and the engine brake state, so the oil pressure change characteristic with respect to the speed ratio e of the hydraulic cylinder changes. For example, FIGS. 7 and 8 show the oil pressure P in the primary hydraulic cylinder 26 in the positive torque state and engine braking state,
, and the oil pressure change characteristics of the oil pressure P0□ in the secondary hydraulic cylinder 28, respectively. u
The magnitude relationship with L is opposite, and in both cases, the hydraulic pressure on the driving side is larger than the hydraulic pressure on the driven side. This phenomenon originally occurred in the primary hydraulic cylinder 26 and secondary hydraulic cylinder 28.
However, in this embodiment, the primary hydraulic cylinder 26 and the secondary hydraulic cylinder 2
Since the pressure receiving areas of 8 are the same, this is directly reflected in the magnitude relationship of the oil pressure.

If it is determined in step S7 that the output torque T is positive, then step s8 is executed, and from the relationship of the following equation (6) stored in advance in the ROM 104, the output torque T is determined to be positive. Thrust force W of the secondary side hydraulic cylinder 28 necessary to generate necessary and sufficient pressure. IIL is calculated based on the actual output torque T of the engine 10 and the actual speed ratio e, and step s9
Then, according to the following equation (7), the above thrust force W0. ．． To the pressure receiving area of the secondary hydraulic cylinder 28, the rotation speed N of the secondary rotating shaft 18. The oil pressure to be supplied to the secondary hydraulic cylinder 28, that is, the second line oil pressure Pi2, is determined based on ut. In addition, in step SIO, from the rotational speed N of the primary rotating wheel 16, n, the increase in oil pressure due to the centrifugal force of the hydraulic fluid in the primary hydraulic cylinder 26 (
Corrected oil pressure) PC is calculated.

Wout=f(To, e)...-
(61 Equation (6) above is calculated in advance to make the tension of the transmission belt 24, that is, the clamping force on the transmission belt 24, a necessary and sufficient value, and the thrust force W o u
t is increased proportionally with the output torque T and the quotient of the speed ratio e. In addition, in the relationship of equation (7), the second term is obtained by subtracting the oil pressure increase due to centrifugal force, which increases with the rotational speed N0□, from the first term to calculate the oil pressure PI! This is for correcting 2. In addition, Cz of (7n formula, (8) formula)
, C+ are centrifugal force correction coefficients, respectively, for the secondary hydraulic cylinder 28. - Determined in advance from the specifications of the next hydraulic cylinder 26 and the specific gravity of the hydraulic oil.

On the other hand, if it is determined in step S7 that the vehicle is in the engine braking state, the direction of power transmission in the belt-type continuously variable transmission 14 is reversed, so steps Sll and S9 are performed which are substantially similar to steps S8 and S9. S
12 is executed, the hydraulic pressure (second line hydraulic pressure) Pffi to be supplied to the primary hydraulic cylinder 26 on the driven side
z is determined. That is, first, in step Sll, the optimum thrust force W in of the primary hydraulic cylinder 26 is calculated based on the output torque T and the speed ratio e from the relationship shown in the following equation (9) stored in advance in the ROM 104, and then In step S12, the thrust W,
The second line oil pressure Plz is determined based on the pressure-receiving area A of the next-side hydraulic cylinder 26 and the rotational speed N in of the next-side rotating shaft 16 . Also, step S
13, the corrected oil pressure Pc is the rotational speed N of the secondary rotating shaft 18 according to the following formula 〇〇. It is calculated from , , .

W,, l= f (T,, e) ・ ・
・(9) Al1゜P c = Cz Nout”
...'αD When the second line oil pressure Pβ2 and the correction oil pressure PC are determined in this way, step S14 is executed next, and the following equation (121
The second line oil pressure control value (2) is determined from the relationship. This is the second line oil pressure Pj! So that z is obtained,
It is determined from a data map stored in consideration of the characteristics of the second pressure regulating valve 58 or an arithmetic expression.

V, = f (PI!z)...
After that, step S15 is executed, and it is determined whether the difference between the actual speed ratio deviation 1e-el/e and the target deviation value ε is smaller than a predetermined constant value. This constant value is set sufficiently small, and the speed ratio deviation l
If the difference between e''-el/e and the target deviation value ε is smaller than a certain value, it means that they are almost defeated. For example, it is a constant value that can be determined through experiments or simulations as a value that provides the best fuel efficiency in the overall driving condition of the vehicle such as driving in 10 modes, or it can be determined as a value that minimizes fuel consumption in various driving conditions. The target deviation value ε may be set by various means, such as setting the minimum value of the speed ratio deviation required to achieve this as the target deviation value ε.

If the judgment in step 315 is No, that is, if the actual speed ratio deviation 1e"-el/e is still not in a state that substantially matches the target deviation value ε, then step 3
16 is executed, and it is determined whether the continuously variable transmission 14 is in a steady state or in a variable speed state. This is, for example, the absolute value of the rate of change of the target speed ratio all or the speed ratio e, 1de"/d
t1. A determination is made based on whether lde/dtl is greater than a certain value, and if it is determined that the steady state is present, steps 317 and 318 are subsequently executed.

These steps S17. S18 is the speed ratio deviation (
In order to determine the first line oil pressure Pi to be controlled by the first pressure regulating valve 48 so that the steady deviation) je''-el/e matches the target deviation value ε, these deviations are fed back as a control amount. In step 517, the proportional action (P action) term F is calculated according to the following formula α■, and in step 818, the integral action (■ action) is calculated according to the following formula α
The term F+ is calculated. Note that KF and Kl are respectively predetermined positive proportionality constants.

・ ・ ・ Circle Once the proportional action term F and the integral action term F1 are determined in this way, step S23 is subsequently executed, and RA
The first line oil pressure (correction value) Pz+' is called from the map stored in M106. This map shows speed ratio e and output torque T. Each time the optimum first line oil pressure Pff is determined by feedback control, step S19. In S20, the oil pressure PR,', which is the sum of the first line oil pressure Pρ1 and the corrected oil pressure P due to the centrifugal force, is stored. Immediately after the start of operation, the optimal first line oil pressure PR, has not yet been determined, so the oil pressure PN, ", which has been stored in advance in the ROM 104, is transferred to the map in the RAM 106.

Then, in the next step S24, the correction oil pressure PC calculated in the step SIO or S13 is subtracted from the called first line oil pressure PR+'' according to the following equation a9, and the proportional pressure PC calculated in the step S17.318 is further subtracted. The operation term FP and the integral operation term F are added to determine the first line oil pressure Pet.In addition, in the subsequent step S25, the first line oil pressure control value is determined from the relationship of the following equation (16) stored in the ROM 104 in advance. ■1 is determined.

Is this the 1st line hydraulic PI? , so that the first
a data map stored in consideration of the characteristics of the pressure regulating valve 48;
Alternatively, it can be obtained from an arithmetic expression. Finally step S2
6 is executed, and the speed ratio control value ■ determined in the previous step together with the first line oil pressure control value ■1.

and the second line oil pressure control value ■2 is output.

Pffl = px, '-PC+FP +l'',
...QS)V, =f(Pf+)
...QS) By repeating these steps, the speed ratio deviation le''-el/e is gradually brought closer to the target deviation value ε, and the first line oil pressure Pβ1 is required to realize the speed ratio e. The pressure is regulated to the minimum required oil pressure.

That is, the output hydraulic pressure characteristics of the speed change control valve 44 are shown in FIG. u
Assuming that a certain speed ratio e is achieved at the oil pressure indicated by O symbol t, the speed ratio control value v0 becomes ΔV0, and this ΔV0 is expressed by the above equation (4), so that the target speed ratio e1 ’ and the speed ratio e is Δ■. A deviation le''-el/e occurs, but this speed ratio deviation 1e''-el/e becomes smaller as the slope of the hydraulic characteristics becomes steeper as the first line oil pressure pH increases. If the first line oil pressure Pit is made smaller, the slope of the oil pressure characteristic becomes gentler and becomes larger. Therefore, if the actual first line oil pressure pH is smaller than the value it should be, the speed ratio deviation le"-el/e is larger, so l
e II-el/e-ε becomes a positive value, and the proportional action term FP and integral action term F also become positive, and the first line oil pressure P
β gradually increases, and the speed ratio deviation 1elI-el
/e is made smaller to finally match the target deviation value ε. In addition, if the actual first line oil pressure Pi is larger than the value it should be, the speed ratio deviation 181''-el/
Since e is small, le"-el/e-ε becomes a negative value, and the proportional action order Fa and integral action term F also become negative, and the first line oil pressure P l + is gradually lowered. , the speed ratio deviation 1e'''-el/e is increased to finally match the target deviation value ε.

Then, when the speed ratio deviation 1e'-el/e and the target deviation value ε are made to match, the actual first line oil pressure realizes the speed ratio e including the target deviation value ε in the operating state at that time. This is the minimum oil pressure value required. That is,
Said steps S17°S18. S23. S24. S25
And S26 functions to feedback-control the first pressure regulating valve 48 so as to regulate the first line oil pressure so that the speed ratio deviation 1e''-el/e matches the target deviation value ε. Note that in this embodiment, the ROM 104 is
Since the hydraulic pressure Fly' stored in is transferred, the amount of correction by feedback control is small, and the speed ratio deviation le"-e1/6 can be made to match the target deviation value ε relatively quickly. .

On the other hand, in this way, the speed ratio deviation 1 e' −e l/
When e and the target deviation value ε are substantially matched and the difference between them is smaller than a certain value, the determination in step S15 becomes YES, and step S19 is executed. In this step 319, the step SIO or S
By adding the corrected oil pressure Pc calculated in step 13,
The first line oil pressure (correction value) PN,' is calculated, and in step S20, which is subsequently executed, the first line oil pressure Pβ,' is calculated in relation to the driving state of the vehicle at that time, and specifically, the output It is stored in a two-dimensional map of torque T8 and speed ratio e.

” + ' = P e + + p c
...αηHere, if you try to store the value of the first line oil pressure pH as it is, the output torque T, (or - next side rotating shaft torque T, , ), speed ratio e, engine rotation speed N
, (or - next-side rotating shaft rotational speed N, 7) must be stored in a three-dimensional map. This is because even if the output torque T and the speed ratio e are the same, the thrust due to the centrifugal force will be different if the rotational speed of the variable boob IJ20.22 is different, and the optimal first line oil pressure Pff.

This is because it also changes. However, such a three-dimensional map is undesirable because it requires a large memory capacity. The oil pressure correction in step S19 was made for this reason, and by adding the corrected oil pressure P corresponding to the thrust due to centrifugal force to the optimal first line oil pressure Pff, the output torque T0 and speed ratio can be adjusted. It was made so that it could be stored in a two-dimensional map with e. Note that the hydraulic cylinder 2 can compensate for the increase in oil pressure due to centrifugal force.
This applies to the oil pressure P iR + P out within 6.28, and cannot be strictly applied to the first line oil pressure PR, which is higher than those oil pressures P i II HP Ou L. Even if the above method is used, the error is considered to be small.

When the first line oil pressure px,' is stored in the two-dimensional map in this way, steps 521 and S22 are executed, and the proportional action term FP of step S17.318 is
, the integral operation 'X F l are both set to 0, and the memory storing the integral value of the integral operation order F1 is also reset. This is the oil pressure P7 corresponding to the optimal first line oil pressure Pβ1! , ' are stored in the two-dimensional map, and the oil pressure Pβ1' is called out in step 323, so correction by feedback is no longer necessary.

Next, step S23. S24 is executed, the first line oil pressure PR+'' is called from the two-dimensional map, and the first line oil pressure pH is determined according to the formula 09. At this time, the proportional action term F2 and the integral action order F
, is 0, so the first line oil pressure Pit is the oil pressure pH
The first line oil pressure PR2 determined in this way is determined by subtracting the corrected oil pressure pc from the speed ratio deviation 1e-el/
e is the optimal oil pressure value that substantially matches the target deviation value ε. Then, by executing steps 325 and 326, the first line oil pressure P/! The first pressure regulating valve 48 is controlled so that .

Further, if it is determined in step S16 that the continuously variable transmission 14 is in a gear change state, then step
21. S22 is executed, and as mentioned above, the proportional action term F
, and the integral operation order F1 are both set to 0, and the memory that stores the integral value of the integral operation order F is also reset. This is because the feedback control is started from the zero state when the steady state is reached thereafter.

Subsequently, steps 323 and subsequent steps are executed, and as described above, the first line oil pressure P! is calculated from the two-dimensional map. 11 is called, and the corrected oil pressure P is calculated from the oil pressure PZl''.

is subtracted, the optimum first line oil pressure P Il + is determined in the operating state at that time, and the first line oil pressure P Il + is determined.
The first pressure regulating valve 48 is controlled so that the line oil pressure PJ is obtained. That is, in this embodiment, the continuously variable transmission 1
4 is in the gear shifting state, the first line oil pressure Pil+
is the 1st line oil pressure PR+” stored in the two-dimensional map.
The hydraulic pressure is controlled moment by moment to achieve the optimum oil pressure value based on the following.

In addition, in the shifting state, the speed ratio deviation 1e''-e1/e
becomes excessive, and the proportional action term F2 and the integral action order F
, becomes very large, so feedback control is not performed during the gear change state.

In this way, even in the shifting state, the first line oil pressure Pff,
is the first line oil pressure Pil stored in the two-dimensional map, '
The first line oil pressure Pβ is stored in the two-dimensional map in step S23 even if the steady state is reached thereafter.

As long as is the optimum value, the first line oil pressure Px is controlled to the optimum oil pressure value such that the speed ratio deviation le"-el/e substantially matches the target deviation value ε. Therefore, even in a steady state, the step If the determination in S15 is YES, no feedback control is performed, and the first line oil pressure P!1 is determined based only on the first line oil pressure Pe1' stored in the two-dimensional map.
Compared to the case where the first line oil pressure Pe1 is gradually corrected by the control, the speed ratio deviation 1 e l''-el/e
The time it takes to match the target deviation value ε is shortened.

Said steps 315, 319. S20. S21.322
．． S23,324. S25. S26 is the speed ratio deviation le
The first line oil pressure PN when ``-el/e substantially matches the target deviation value ε is stored in association with the vehicle operating condition, and feedback control is performed when the same operating condition occurs thereafter. The first pressure regulating valve 48 is controlled based on the stored first line oil pressure Pi,'.

However, the first line oil pressure PI stored in the two-dimensional map
! , is inappropriate and the first line oil pressure Pf is determined based only on the first line oil pressure pH, and if the speed ratio deviation 1e''-el/e does not match the target deviation value ε, The judgment in step 515 is NO, and feedback control is performed. Then, when the speed ratio deviation 1e - e l / e is made to substantially match the target deviation value ε by this feedbank control, step S19.32
0 is executed, and a new first line oil pressure PIl+' calculated by adding the corrected oil pressure Pc to the first line oil pressure P R+ at that time is stored in the two-dimensional map.

For this reason, for example, if there is a change in the hydraulic characteristics of the speed change control valve 44 due to a difference in the viscosity of the hydraulic oil when it is cold immediately after the start of operation and after it has been sufficiently warmed up.

If there is a change over time in the speed change control valve 44 or the first pressure regulating valve 48 and the pressure regulation accuracy changes, or for some other reason an abnormal value may appear in the first line oil pressure P! 1. Even in the case where the first line oil pressure Pl, ° is stored separately, the first line oil pressure Pl, ° is sequentially rewritten to the optimum value for the operating state, and the first line oil pressure Pβ1 determined based on the first line oil pressure Pff, °. is always controlled to the minimum oil pressure necessary to make the speed ratio deviation (e''-el/e coincide with the target deviation value ε).

In this way, in this embodiment, the speed ratio deviation 1e''-
The first pressure regulating valve 4 is adjusted so that el/e matches the target deviation value ε.
8 is feedback-controlled, the actual first line oil pressure regulated by the first UEI pressure valve 48 is the minimum oil pressure value necessary to realize the speed ratio e including the target deviation value ε, and the pump The driving loss of the engine 42 and further the power loss of the engine 10 are reduced, and the fuel efficiency of the vehicle is improved. In particular, in this example, the target deviation value ε is
Since the value is set to minimize the fuel efficiency of the vehicle, the fuel efficiency of the vehicle is significantly improved.

In addition, since the first line oil pressure is regulated based on the speed ratio deviation 1 e'' -e l/e, the first line oil pressure is adjusted in the pressure regulating system including the first pressure regulating valve 48 and the second pressure regulating valve 58 Even if there is a pressure adjustment error in the oil pressure or the second line oil pressure, or if there are variations in the characteristics of the speed change control valve 44,
The actual first line oil pressure is always controlled to the minimum necessary oil pressure value. That is, as long as a certain correlation is ensured between the first line oil pressure control value (1) and the actual first line oil pressure, the absolute value of the oil pressure does not need to be very accurate. Therefore, the pressure regulating system and the speed change control valve 44
Therefore, it is not necessary to use a highly accurate one, and the manufacturing cost can be reduced.

On the other hand, due to the above feedback control all, the speed ratio deviation is 1
When eゝ-el/e is made to substantially match the target deviation value ε, the oil pressure PJ, which is the sum of the correction oil pressure PC and the first line oil pressure P1+ at that time, is calculated as the output torque T and the speed ratio e. It is stored in a two-dimensional map, and when the same operating condition occurs thereafter, the first line oil pressure Pl is determined based on the stored oil pressure pH'.

Since it is designed to determine the first line oil pressure P6
．． The pressure is immediately regulated to the optimum oil pressure value that realizes the speed ratio e including the target deviation value ε, and a more efficient reduction in fuel consumption can be achieved compared to the case where the pressure is regulated by constant feedback control.

In particular, in this embodiment, even when the continuously variable transmission 14 is in a gear change state, the first line oil pressure PIl+ is the oil pressure Pj7. Since the oil pressure is controlled to the optimum oil pressure value based on the above, there is an advantage that the drive loss of the pump 42 during gear change is significantly reduced.

Incidentally, Fig. 9 shows that when the throttle valve opening θ is decreased,
In other words, the first
Line oil pressure P6. In the figure, the solid line is based on the present example, and the dashed line is the change in PI based on the speed ratio deviation.
In the case of feedback control by operation, the two-dot chain line shows the case where the addition function of the integral action term F is stopped and the feedback action is primarily limited by reducing the gain of the proportional action term F2. As is clear from FIG. 9, in this embodiment, the oil pressure is controlled to the minimum necessary value at each speed ratio during gear shifting, but when feedback control is used, the oil pressure becomes more than necessary, causing the pump 42 to be driven. This results in losses.

In addition, in this embodiment, the first line oil pressure P determined based on the first line oil pressure PR8" stored in the two-dimensional map
For R+, speed ratio deviation]eゞ-e]/e and target deviation value ε
If they do not match, the first line oil pressure P R+
Since ° is rewritten sequentially, even if the pressure regulation accuracy of the first pressure regulating valve 48 changes, etc.
There is an advantage that the optimum first line hydraulic pH is always achieved immediately.

Note that in this embodiment, among the series of signal processing by the controller 94, step S15. S17. S18, S19.
S20. S21. S22. S23°S24. S25 and S26 correspond to control means for controlling the first pressure regulating valve 48.

On the other hand, conventionally, for example, in the case of a positive torque state, the first line oil pressure Pβ1 was determined according to the following equations (18) to Qυ. That is, first, the target speed ratio e1 and the engine 1
The thrust ratio γ is based on the output torque T, which is 0. (Thrust force W of the secondary hydraulic cylinder 28. uL/- the secondary hydraulic cylinder 26
The thrust force W, 7) is calculated, and the thrust ratio T and the thrust force W of the secondary hydraulic cylinder 28 are calculated from equation 01. , the thrust force W in of the next hydraulic cylinder 26 is determined. Next, from the Q@ equation, the thrust force W of the primary side hydraulic cylinder 26, 7.
- Pressure receiving area A of the next side hydraulic cylinder 26, 7. - Based on the rotational speed N in of the next rotating wheel 160, the target speed ratio e
Calculate the oil pressure p required for the primary side hydraulic cylinder 26 to realize 1, and further calculate the oil pressure P1 according to formula 09.
The first line oil pressure Pff was determined by adding the excess oil pressure ΔP1 to , .

r, =f (e”, T, )...
-QBlHere, the above-mentioned surplus oil pressure ΔP1 of Qυ2 is necessary for reducing the steady deviation le"-el/e of the speed ratio. That is, in this embodiment, the above-mentioned fourth
As shown in the figure, a steady deviation 1e"'-el/e corresponding to ΔV0 always occurs, but this steady deviation 1e"-el/e can be changed by increasing the first line oil pressure PJ. Since the slope of the characteristic becomes steeper, the first line oil pressure PN becomes smaller.

If it is made smaller, the slope of the hydraulic characteristics becomes gentler, so it becomes larger. However, as the first line oil pressure PJ is increased, the driving loss of the pump 42 also increases accordingly, so the margin oil pressure ΔP1 is determined at a balance point between the driving loss and the steady-state deviation, which are contradictory to each other.

On the other hand, the oil pressure P obtained from the above aω2 to Qω2 is 0
may not necessarily be completely different from the actual oil pressure of the primary side hydraulic cylinder 26 at the target speed ratio e8 due to individual differences in each part constituting the continuously variable transmission 14, changes over time, pressure adjustment errors in the first pressure regulating valve 48, etc. It doesn't necessarily match. therefore,
If we try to control the steady-state deviation le"-el/e so that it is always smaller than a certain value, the hydraulic pressure P calculated above
, , pressure adjustment error of the first pressure regulating valve 48, etc., it was necessary to set the excess oil pressure ΔP1 to a large value.

For this reason, an unnecessarily high first line oil pressure Pβ1 is prepared in an operating range with few errors or in a period where no changes occur over time, causing a driving loss of the pump 42 and furthermore a power loss of the engine 10. The vehicle's fuel efficiency was impaired.

Although one embodiment of the present invention has been described above in detail based on the drawings, the present invention can also be implemented in other embodiments.

For example, in the embodiment described above, if the optimum first line oil pressure P/, is not yet achieved immediately after the start of operation, the oil pressure PN,' stored in advance in the ROM 104 is stored in the RAM.
Although it is now transferred to the two-dimensional map of 106, as in the past, the first line oil pressure (calculated value) px determined in 2112 is transferred to the U expression.

It is also possible to determine the first line oil pressure PI using only feedback control. When determining only by feedback control, it is desirable to add a differential action (D action) term to prevent hunting.

In addition, although the feed bank control in the above embodiment determines the first line oil pressure P 6 + by the so-called PI operation, it is also possible to perform control based only on either the P operation or the ■ operation. be.

Further, in the above embodiment, the deviation l ell-
Although elle is used, it is also possible to perform feedbank control using le'-el as the control amount.

Furthermore, in the embodiment described above, the first line oil pressure PA,' stored in the map is sequentially rewritten to the optimum oil pressure, but once the first line oil pressure P is set to the optimum oil pressure! ! , ' is stored, the hydraulic pressure PI! is always stored without rewriting the value. , "The first line oil pressure Pf may be determined based on ".

Furthermore, in the above embodiment, the hydraulic pressure Pg, which is the sum of the first line hydraulic pressure P18 and the corrected hydraulic pressure P due to centrifugal force, is stored in the two-dimensional map of the output torque Te and the speed ratio e. Torque Te.

It is also possible to store the first line oil pressure PR+ as it is in a three-dimensional map or the like consisting of the speed ratio e and the engine rotational speed N3.

Further, in the above embodiment, the target deviation value ε is set to a constant value that minimizes the fuel efficiency of the vehicle, but this target deviation value ε is set as appropriate depending on the purpose, and may also be set depending on the driving of the vehicle. Data map the optimal target deviation value ε according to the condition,
It is also possible to obtain it using an arithmetic expression or the like.

Furthermore, in the embodiment described above, the feedback ill? Since there is a constant relationship between and the rotational speed N of the primary rotating shaft 16, the steady-state deviation IN, -Ni, between the rotational speed N in and the target rotational speed N i n''
It is also possible to make the speed ratio deviation 1eゝ-elle coincide with the target deviation value ε by performing feedback control based on l/Ni, .

Although other examples are not given, the present invention can be implemented with various modifications and improvements based on the knowledge of those skilled in the art without departing from the spirit thereof.

[Brief explanation of the drawing]

FIG. 1 is a diagram showing the configuration of a hydraulic control system for a vehicle heald-type continuously variable transmission, which is an embodiment of the present invention. 2 and 3 are flowcharts for explaining the operation of the embodiment of FIG. 1. FIG. 4 is a diagram showing the output oil pressure characteristics of the speed change control valve of FIG. 1. FIG. 5 is a diagram showing the relationship between the throttle valve opening degree and the target rotational speed of the primary rotating shaft in the embodiment of FIG. 1. FIG. 6 is a diagram showing the minimum fuel consumption rate curve of the engine of FIG. 1. Figures 7 and 8 are
FIG. 7 is a diagram showing the change characteristics of the oil pressure of each part with respect to the speed ratio in the embodiment shown in the figure, and FIG. 7 shows the positive torque state,
FIG. 8 shows the engine braking condition. FIG. 9 is a diagram showing the change in the first line oil pressure with respect to the speed ratio change in the embodiment of FIG. 1, and is also a diagram showing the case where only feedback control is used. 14: Herto 2 continuously variable transmission 16 Secondary rotating shaft 18: Secondary rotating shaft 20 2 Primary variable pulley 22: Secondary variable pulley 24: Transmission belt 26 Secondary hydraulic cylinder 28: Secondary Side hydraulic cylinder 44: Speed control valve 48: First pressure regulating valve 58: Second pressure regulating valve 94: Controller Applicant: Toyota Motor Corporation Fig. 4 (+o3rpm) Fig. 5 L rotation a-side Fig. 7 Fig. 8 Figure 9 Sun-F Run n

Claims (2)

[Claims] (1) A pair of primary variable pulleys and a secondary variable pulley provided on the secondary rotating shaft and the secondary rotating shaft, respectively, and a transmission belt that is wound around the pair of variable pulleys to transmit power. , a belt-type continuously variable transmission for a vehicle comprising a pair of primary hydraulic cylinders and secondary hydraulic cylinders that respectively change the effective diameters of the pair of variable pulleys, wherein the pressure of hydraulic oil supplied from a hydraulic source is adjusted. a first pressure regulating valve that supplies the hydraulic oil regulated to the first line hydraulic pressure to one of the primary hydraulic cylinder and the secondary hydraulic cylinder, and simultaneously controls the hydraulic oil in the other cylinder; A speed change control valve that adjusts the speed ratio of the continuously variable transmission by changing the effective diameters of the primary variable pulley and the secondary variable pulley by causing the flow to flow from the primary hydraulic cylinder and a second pressure regulating valve that regulates the pressure of the hydraulic oil flowing out from the other side of the secondary hydraulic cylinder to a second line hydraulic pressure lower than the first line hydraulic pressure, and the actual speed ratio is adjusted according to the driving state of the vehicle. A hydraulic control device of a type that controls the speed change control valve so as to match a target speed ratio determined according to the above, wherein a steady deviation between the actual speed ratio and the target speed ratio is a predetermined target deviation value. The first pressure regulating valve is feedback-controlled so as to regulate the first line hydraulic pressure so as to match the target deviation value.
Control means for storing the line oil pressure in relation to the operating state of the vehicle, and controlling the first pressure regulating valve based on the stored value without performing the feedback control when the same operating state subsequently occurs. A hydraulic control device for a belt-type continuously variable transmission for a vehicle, comprising: (2) The control means adds an increase in oil pressure due to centrifugal force to the first line oil pressure and stores the result in a two-dimensional map of the speed ratio and engine output torque. A hydraulic control device for the vehicle belt-type continuously variable transmission described above.