JPS63152756A - Hydraulic control for vehicle belt type continuously variable transmission - Google Patents

Abstract

PURPOSE:To simplify the structure of the control in the caption and improve its reliability, by feeding discharged oil from a pump to the primary and secondary side pulleys via a distributing valve and providing each feed piping with the 1st and the 2nd pressure regulating valves. CONSTITUTION:An oil pump 42 is connected to the primary side hydraulic cylinder 26 of the primary side variable pulley 20 and the secondary side hydraulic cylinder 28 of the secondary side variable pulley 22, via a distributing valve 44, the 1st line oil passage 62, and the 2nd line oil passage 64, with the 1st line oil passage 62 and the 2nd line oil passage 64 provided with the 1st pressure regulating valve 68 and the 2nd pressure regulating valve 70 respectively. Hence, transmission control can be performed by simply regulating the 1st line oil pressure Pl1 and 2nd line oil pressure Pl2, then no transmission control valve is required, the structure of a control is simplified and the reliability is improved.

Description

DETAILED DESCRIPTION OF THE INVENTION Technical Field The present invention relates to an improvement in a hydraulic control device for a belt-type continuously variable transmission for a vehicle.

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.

In the hydraulic control device for such a continuously variable transmission, for example,
As described in Publication No. 2-98861, the tension of the transmission belt is adjusted exclusively by adjusting the working pressure supplied to the secondary side (driven side) hydraulic cylinder, and The speed ratio is controlled exclusively by adjusting the flow rate of hydraulic fluid supplied to or discharged from the hydraulic fluid.

In the hydraulic control system of this type of continuously variable transmission, one type of line hydraulic pressure is prepared that is regulated in relation to the actual speed ratio, etc., and this is sent exclusively to the secondary hydraulic cylinder that maintains the tension of the transmission belt. At the same time, it is supplied to the primary hydraulic cylinder via a flow control valve that controls the speed ratio. Therefore, the flow rate of the hydraulic oil supplied to the primary hydraulic cylinder or the hydraulic oil discharged from it changes depending on the line oil pressure, that is, the speed ratio of the continuously variable transmission, etc. It is unavoidable that it will be influenced by the speed ratio etc.

Therefore, sufficient transient response may not be obtained in controlling the speed ratio. Furthermore, when the direction of power transmission is reversed during engine braking of a vehicle, in effect, the driven side hydraulic cylinder exclusively controls the speed ratio, and the driving side hydraulic cylinder controls the transmission belt tension. Therefore, there was a drawback that the control characteristics of the tension and speed ratio of the transmission belt could not be obtained properly.

On the other hand, as described in Japanese Patent Publication No. 58-29424, there is a control valve (four-way control valve) that changes the speed ratio by simultaneously supplying hydraulic oil from a hydraulic source to one side of the hydraulic cylinder and letting it flow out from the other side. A hydraulic control device for a belt-type continuously variable transmission is provided, which is equipped with a pressure regulating valve that regulates the pressure of hydraulic oil flowing out from the control valve. In this type of hydraulic control device, a relatively high hydraulic pressure from a hydraulic source is applied to the hydraulic cylinder located on the side (drive side) where the internal hydraulic pressure is high in the power transmission state among the two hydraulic cylinders, Since the hydraulic pressure regulated by the pressure regulating valve is applied to the hydraulic cylinder on the opposite side, the tension and speed ratio of the transmission belt can be suitably controlled even if the power transmission direction is reversed. However, in this type of hydraulic control device, the pressure of the hydraulic source is not controlled and is only maintained at a constant pressure by a normal relief valve. If the oil pressure value in the engine changes, the speed ratio change speed, that is, the speed change responsiveness may not be constant. Also,
On the other hand, if the pressure of the hydraulic power source is set high in anticipation of a large margin of oil pressure in order to obtain a sufficient rate of speed ratio change over the entire operating condition, there is a drawback that the power loss required to constantly maintain that pressure becomes large. Ta.

In response, the present applicant created a hydraulic control device to solve this problem and filed an application earlier.

This is the hydraulic control device described in Japanese Patent Application No. 61-37571. This hydraulic control device includes (1) a first pressure regulating valve that regulates the pressure of hydraulic oil supplied from a hydraulic source to obtain a first line hydraulic pressure; and (2) a first pressure regulating valve that regulates the pressure of hydraulic oil supplied from a hydraulic source to a first line hydraulic pressure; By simultaneously supplying hydraulic oil to one of the side hydraulic cylinder and the secondary side hydraulic cylinder and flowing out the hydraulic oil in the other side, the effective diameters of the primary side variable boob and the secondary side variable pulley are changed, thereby controlling the continuously variable transmission. (3) adjusts the pressure of the hydraulic oil flowing out from the other of the primary hydraulic cylinder and the secondary hydraulic cylinder through the transmission control valve, and adjusts the pressure of the hydraulic oil flowing out from the other of the primary hydraulic cylinder and the secondary hydraulic cylinder, and 2
It is configured to include a second vI4 pressure valve that uses line oil pressure. In the hydraulic control device configured in this way, the first line hydraulic pressure and the second line hydraulic pressure are prepared by the first pressure regulating valve and the 21st J8 pressure valve, so the pressure difference between them allows the primary side hydraulic cylinder and the secondary side hydraulic pressure to be adjusted. The flow rate of hydraulic oil supplied to or discharged from one of the cylinders is determined. Therefore, the speed ratio change speed is determined according to the differential pressure between the first line oil pressure and the second line oil pressure regardless of the speed ratio of the continuously variable transmission, so that sufficient transient response of speed ratio control can be obtained. Furthermore, by controlling the first pressure regulating valve in relation to the output state of the engine, the first line oil pressure is controlled to a necessary and sufficient value so that a sufficient speed ratio change rate is obtained and no power loss occurs. At the same time, by controlling the 2nd FI pressure valve in relation to the speed ratio and transmission torque, the 2nd line oil pressure is controlled to a necessary and sufficient value within a range that does not cause slippage of the transmission belt, so the power loss of the vehicle is significantly reduced. It has the advantage of being reduced.

Problems to be Solved by the Invention By the way, the speed change control valve of the above-mentioned hydraulic control device operates two types of first line hydraulic pressure and second line hydraulic pressure reciprocally on the primary hydraulic cylinder and the secondary hydraulic cylinder to adjust the speed ratio. be kept constant or changed. For example, when the hydraulic pressure P1 in the primary hydraulic cylinder is larger than the hydraulic pressure PomL in the secondary hydraulic cylinder at a speed ratio in a predetermined operating state, or when performing an increasing speed change, the shift control valve The first line hydraulic pressure is applied to the primary hydraulic cylinder, and at the same time, the second line hydraulic pressure is applied to the secondary hydraulic cylinder. On the other hand, when P5 is smaller than PouL at the speed ratio, or when a deceleration shift is performed, the shift control valve applies the first line hydraulic pressure to the secondary hydraulic cylinder and Apply 2-line hydraulic pressure to the primary hydraulic cylinder. In this way, the speed change control valve not only controls the flow rate by changing its flow cross-sectional area (valve opening degree), but also has the control function as a switching valve by switching the hydraulic cylinder to which two types of line hydraulic pressure are applied. ing. For this reason, a spool valve element is provided which is positioned at a neutral position by two springs and moved left and right by two solenoids.

In addition, patent application No. 61-37575 previously filed by the present applicant
As described in the above issue, instead of the above-mentioned type of speed change control valve, it is also considered to install two sets of flow rate control valves to provide the same function.

Therefore, according to the conventional hydraulic control device, two sets of springs and solenoids are required to hold the spool valve in the neutral position, or two sets of flow control valves are required. As a result, the structure becomes complicated and the number of component parts increases, making it difficult to obtain sufficient reliability and increasing the cost.

Means for Solving the Problems The present invention has been made against the background of the above circumstances.
The gist is that a pair of primary variable pulleys and a secondary variable pulley are provided on the primary rotating shaft and secondary rotating shaft, respectively, and that power is transmitted by being wrapped around these pair of variable pulleys. A hydraulic control device for a belt-type continuously variable transmission for a vehicle, comprising a power transmission belt, and a pair of primary and secondary hydraulic cylinders that respectively change the driven diameters of the pair of variable pulleys, al comprises a first valve chamber and a second valve chamber that allow hydraulic oil to flow out to a pair of first output port and second output port, respectively, and a predetermined constant difference between the first valve chamber and the second valve chamber; a distribution valve for distributing hydraulic fluid pumped from a hydraulic source to flow out from a first output port and a second output port so as to generate pressure; A first line hydraulic pressure and a second line are connected to the ports, respectively, for controlling the pressure of the hydraulic oil flowing out from the first output port and the second output port to obtain a target speed ratio and maintain the tension of the transmission belt at an appropriate level. It includes a pair of first pressure regulating valve and second pressure regulating valve that regulate the hydraulic pressure.

Operation and Effects of the Invention With this arrangement, unlike conventional hydraulic control devices, it is only necessary to control the first line hydraulic pressure and the second line hydraulic pressure, eliminating the need for a speed change control valve, simplifying the structure, and increasing reliability. Improve customer service. In addition, the first pressure regulating valve exclusively controls the hydraulic pressure of the primary side hydraulic cylinder, and the second pressure regulating valve exclusively controls the hydraulic pressure of the secondary side hydraulic cylinder, but the working pressure after a differential pressure is established in advance by the distribution valve. It is only necessary to adjust the pressure against
Since pressure control valves having exactly the same specifications can be used for both pressure regulating valves, manufacturing of the device becomes easy and the device becomes inexpensive.

Here, one of the pair of pressure regulating valves preferably transmits the working hydraulic pressure regulated to the first line hydraulic pressure or the second line hydraulic pressure to the primary side hydraulic cylinder or the secondary side hydraulic cylinder via a flow rate control valve. supply In this case, if the flow control valve is operated in a direction that reduces the deviation between the target speed ratio and the actual speed ratio, more accurate and stable speed ratio control becomes possible. The flow control valve is composed of a set of simple valves capable of controlling the flow rate, - the oil pressure control device is simple and inexpensive.

Further, the flow rate control valve is preferably provided in an oil passage that supplies the first line hydraulic pressure regulated by the first pressure regulating valve to the primary hydraulic cylinder.

Preferably, one of the first pressure regulating valve and the second pressure regulating valve transmits the relatively lower hydraulic pressure of the first line hydraulic pressure and the second line hydraulic pressure to the torque transmission of the belt type continuously variable transmission. The pressure is regulated to a value that generates a necessary and sufficient tension in the transmission belt. The hydraulic pressure on the side is increased to 81iI immediately to generate the thrust necessary to maintain the speed ratio of the belt type continuously variable transmission at the target speed ratio.

Preferably, the flow rate control valve increases a flow cross-sectional area in response to a deviation between an actual speed ratio and a target speed ratio of the belt type continuously variable transmission, and 1
The second pressure regulating valve of the ill valve pressure valve increases the oil pressure of the pair of primary variable pulley and secondary variable pulley having a relatively smaller thrust, or This is to lower the oil pressure of the primary variable pulley and the secondary variable pulley that have a relatively large thrust.

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 rotating wheel 16, a secondary rotating shaft 18, and a primary variable pulley 20 with a variable effective diameter attached to the primary rotating shaft 16 and the secondary rotating shaft 18. and the secondary variable boolean I722, the transmission belt 24 wound around the primary variable pulley 20 and the secondary variable pulley 22, the primary variable pulley 20 and the secondary variable boolean I
It includes a primary hydraulic cylinder 26 and a secondary hydraulic cylinder 28 that change the effective diameter of J22. The primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28 are formed to have the same pressure receiving area, and the primary variable pulley 20 and the secondary variable pulley 22 have the same external shape, so that the belt-type The gear transmission 14 is made small.

The primary side variable pulley 20 and the secondary side variable boob IJ22 are connected to the primary side rotating wheel 16 and the secondary side rotating wheel 1.
fixed rotating bodies 30 and 32 respectively fixed to 8;
Movable rotary bodies 34 and 36 are provided on the primary rotary ring 16 and the secondary rotary shaft 18 so as to be non-rotatable and movable in the axial direction, respectively, and form a V-groove with the fixed rotary bodies 30 and 32. It consists of

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, an operating 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 oil pump 42 and fed under pressure to the input port 46 of the distribution valve 44. This oil pump 42 is
It constitutes the hydraulic power source of this embodiment, and is driven by the engine 10 via a drive shaft (not shown).

The distribution valve 44 has a cylindrical bore 51 through which the input port 46, the first output port 48, and the second output port 50 communicate with each other, and is slidably fitted into the bore 51 so that the first output port 48, and a second valve chamber 54, which communicates with the second output port 50.
A valve element 56 that reciprocally changes the flow cross-sectional area of the valve 50, and a spring 58 that biases the valve element 56 in the direction of opening the first output port 48 and closing the second output port 50 are provided. The valve element 56 has a cylindrical shape, and its longitudinal hole 59 is closed at the middle part in the longitudinal direction by a partition wall 61 having a crest hole 60.
and the second valve chamber 54 are communicated with each other. Therefore, the pressure receiving area of the left and right end surfaces of the valve element 56 is S, the working pressure in the first valve chamber 52 is PI, the working pressure in the second valve chamber 54 is Pt, and the biasing force of the spring 58 is F (strictly speaking, Benko 56
(which is considered to be approximately constant), the valve element 56 is positioned according to the following equation (11).

P,・S +F=Pz・S...(1)P
2 = PI + F/S ... (2) Above (
As shown in equation (2) converted from equation 11, in the distribution valve 44, the working pressure Pt in the second valve chamber 54 is lower than the first valve chamber 5.
In other words, the pressure loss Δ is constant at the hole 60 so that the hydraulic pressure P1 in the hole 60 is always higher by a constant value F/S.
The hydraulic oil supplied to the input port 46 is distributed so that p (=pz PI) is formed, and the hydraulic oil is distributed to the first output port 4.
8 and the second output port 50 are adjusted. The first line oil passage 62 is provided with a first pressure regulating valve 68 that regulates the first line oil pressure PII by changing the release amount of hydraulic oil in response to a command from the control device 66. The passage 64 is provided with a second pressure regulating valve 70 that regulates the second line hydraulic pressure PR by changing the amount of hydraulic oil released in response to a command from the control device 66. Working oil pressure P+ in the first valve chamber 52 and the second valve chamber 54
A constant pressure difference ΔP is always formed between the hydraulic pressure Pt and the hydraulic pressure Pt in the line, regardless of the values of the first line hydraulic pressure PR and the second line hydraulic pressure PR. For example, first line oil pressure P6
．． When the second line oil pressure P12 is set higher than
When the flow rate from 4 to the drain is restricted, the flow rate passing through the crest hole 60 increases, the pressure loss generated therein increases, and the valve element 56 is moved against the biasing force of the spring 58.

Therefore, the outflow amount at the first output port 48 is reduced until the pressure loss reaches the constant differential pressure ΔP. On the other hand, the first line oil pressure Pn is the second line oil pressure PI2.
In the case where the flow rate is set higher than that, when the flow rate from the first line oil passage 62 to the drain is throttled by the operation of the first pressure regulating valve 68, the flow rate passing through the crest hole 60 decreases, and the pressure loss generated there. becomes smaller and the valve 56 becomes the spring 58.
It is moved according to the urging force of. Therefore, the outflow amount at the second output port 50 is reduced until the pressure loss reaches the constant differential pressure ΔP. In such an operation, the flow rate passing through the crest hole 60, in other words, the flow rate from the first output port 48 is controlled to be constant, so that the first line oil pressure P1. and second line oil pressure P
18, the differential pressure between them is maintained.

The hydraulic oil in the first line oil passage 62 whose pressure is regulated by the first pressure regulating valve 68 is transferred to a flow control valve 72 provided in the primary oil passage 71.
is guided to the primary side hydraulic cylinder 26 via
The hydraulic oil in the second line oil passage 64 whose pressure is regulated by the second pressure regulating valve 73 is guided to the secondary hydraulic cylinder 28 via the secondary oil passage 73.

The flow rate control valve 72 is a two-bottom, two-position type valve,
The input port 75 and the output port door 6 connected to the first pressure regulating valve 68, a spool valve element (not shown) that continuously changes the flow cross-sectional area between them, and the spool valve element is biased in the valve closing direction. and an electromagnetic solenoid 84 that drives the spool valve element. Therefore, when the electromagnetic solenoid 84 is excited, the input port door 5 and the output port door 6 are opened.
When the electromagnetic force is weakened as the excitation state of the electromagnetic solenoid 84 is weakened, the flow cross-sectional area between the input port door 5 and the output port door 6 is continuously reduced. That is, the flow rate control valve 72 functions as a variable throttle, thereby changing the speed ratio change speed of the belt type continuously variable transmission 14.

In a vehicle, the rotation speeds N1 and N of the primary rotating shaft 16 and the secondary rotating wheel 18 of the belt type continuously variable transmission 14 are
A first rotation sensor 86 and a second rotation sensor 88 for detecting 0□ are provided near the primary variable pulley 20 and the secondary variable pulley 22, and the first rotation sensor 86 and the second rotation sensor From 88, signals SRI and SR2 corresponding to the rotational speeds of the primary rotating wheel 16 and the secondary rotating wheel 18 are supplied to the control device 66.

Further, a throttle sensor 92 is provided in the intake pipe of the engine 10 to detect the opening degree θ of the stroke valve 90, and a signal Sθ representing the stroke valve opening degree θ is sent from the throttle sensor 92 to the control device. 66. Further, the engine 10 is provided with an engine rotation sensor 94 for detecting the engine rotation speed N, and the engine rotation speed N is detected from the engine rotation sensor 94.

A signal SE representing .

The control device 66 is a so-called microcomputer including a CPU, ROMSRAM, etc., and the CPU is a RAM.
The first pressure regulating valve 68,
Drive signals PDI, FD2, and FD for controlling the second iPl pressure valve 70 and flow rate control valve 72 are output.

The operation of this embodiment will be explained below with reference to the flowchart in FIG.

First, by executing step S1, the rotational speed N! of the primary rotating wheel 16 is increased! The rotation speed N0Ilt of the R1 secondary rotating wheel 18, the opening degree θ1 of the stroke valve 90, and the engine rotation speed N are read based on the input signal. Next, in step S2, the following equation (
According to 3), the actual speed ratio e of the belt type continuously variable transmission 14
is calculated.

e=N, uL/N 7 (3) In the following step S3, the target rotational speed N in '' is calculated based on the throttle valve opening θ from the relationship previously stored in the ROM. The relationship (Ni-" = f (θ)) for determining the target rotational speed N,- is shown, for example, in FIG. In addition, in step S4, the target rotational speed N, - and the rotational speed N of the secondary rotating wheel 18 are determined based on the relationship shown in equation (3) above. A speed ratio eI' is calculated.

Then, in step S5, the control amount ■, which is the content of the drive signal FD for the flow rate control valve 72, is determined based on the following equation (4).
. is calculated. In step S16, which will be described later, the hydraulic oil flow rate is adjusted based on this control amount v0. That is,
As the deviation between the target speed ratio e''' and the actual speed ratio e becomes larger, the flow rate of the hydraulic oil is increased to improve responsiveness and stability. Note that k in equation (4) is a constant.

Vo = k ・l e” −e l/e ・・14
) Then, in step S6, the actual torque T0 of the engine 10 is determined by a well-known function (T.

=f(θ, N,)) is calculated based on the actual throttle valve opening θ and the engine rotational speed N. The above functions are stored in ROM in advance. Then, in step S7, it is determined whether the vehicle is in a positive drive state or an engine braking state according to the sign of the actual torque T determined as described above.

The reason why such a judgment is necessary is that if the power transmission direction differs between the positive torque state and the engine brake state, the oil pressure P in the hydraulic cylinders 26 and 28 will change. M and P, □
This is because the oil pressure change characteristics with respect to the speed ratio e are different.

For example, FIGS. 5 and 6 show the oil pressure P in the primary hydraulic cylinder 26 and the oil pressure P in the secondary hydraulic cylinder 28 in the positive torque state and the engine brake state.
The oil pressure change characteristics of out are shown, and the oil pressure P
! 7 and the hydraulic pressure P out are opposite in magnitude, and in both cases, the hydraulic pressure on the drive side is larger than the hydraulic pressure on the driven side. This phenomenon is originally discussed based on the thrust forces of the primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28, but in this embodiment, the pressure receiving area of the primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28 is Since they are equivalent,
This is directly reflected in the magnitude relationship of hydraulic pressure.

If it is determined that the output torque T is positive in step S7, the vehicle is being driven in the normal direction, and the flow rate control valve 72
The output characteristic of is shown in FIG. If it is determined that the drive is normal as described above, in step S8,
Thrust force necessary for torque transmission, in other words, thrust force W for generating necessary and sufficient narrowing pressure on the transmission belt 24
A second line hydraulic pressure Pit for generating o u L in the secondary hydraulic cylinder 28 is determined. That is, first, the optimal thrust of the secondary hydraulic cylinder 28 (
Calculated value) Wout° is calculated. Also, the following formula (
6), the above thrust W0□'', secondary hydraulic cylinder 2
8 pressure receiving area A0□, rotation speed N of the secondary rotating wheel 18
Hydraulic pressure (calculated value) based on 6 ut Pout.

is calculated. This value P0□ ” is the thrust force W.
This is the value necessary to generate. Then, according to the following equation (7) stored in advance in the ROM, the second line oil pressure P1 is calculated from P out ' determined by equation (6).
2 is calculated. In this case, since the second line oil passage 64 is directly connected to the secondary hydraulic cylinder 28 by the secondary oil passage 73, the hydraulic pressure correction value ΔP is zero, and Pit and P out' are the same. becomes.

WotIt 1=f (To, e) ...(
5) Here, the above equation (5) is calculated in advance in order 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 Wou
t9 is increased proportionally with output torque T0 and speed ratio e. Also, in the relationship of equation (6),
The second term is the rotation speed N. This is to correct the oil pressure P outo by subtracting the centrifugal oil pressure that increases with U from the first term. The second term C1 is a centrifugal force correction coefficient, which is determined in advance from the specifications of the secondary hydraulic cylinder 28 and the specific gravity of the hydraulic oil.

In the subsequent step S9, the oil pressure (target oil pressure) Pi in the primary side hydraulic cylinder 26 is determined in order to generate necessary and sufficient thrust to realize the target speed ratio e1. That is, first, based on the target speed ratio eI' and the actual output torque T of the engine 10, from the relationship shown in the following equation (8) stored in advance in the ROM, the thrust ratio T during forward drive, (secondary side hydraulic pressure The thrust force W0□ of the cylinder 28/the thrust force W i - of the primary hydraulic cylinder 26 is calculated, and the thrust ratio γ and the thrust force W6 of the secondary hydraulic cylinder 28 are calculated from the following equation (9). ．． The thrust force Wiat of the primary hydraulic cylinder 26 is determined from L'. And the following formula 0
1 to the thrust force W0 of the primary hydraulic cylinder 26, the pressure receiving area A!R1 of the primary hydraulic cylinder 26, the rotational speed N!R of the primary rotating shaft 16, and calculate the hydraulic pressure (calculated value) P!n'. , From the following formula 〇〇, the above oil pressure p , , % and based on the corrected oil pressure ΔP1, the primary side inlet oil pressure Pl,
is calculated.

γ. =r (e”, Ts)...1
8) Here, the above equation (8) shows a relationship determined in advance so that the thrust ratio γ for obtaining the target speed ratio e1 can basically be determined over a wide range of operating conditions, and The first line oil pressure is controlled so that the thrust ratio T determined in relation to the target speed ratio e8 and the actual output torque T is obtained from this relationship. In addition, in the relationship of the above α0 equation, the second term is corrected by subtracting the centrifugal oil pressure that increases with the rotational speed N in from the first term, and C2 of the second term is a correction for the primary side hydraulic cylinder 26.
It is determined in advance from the specifications of and the specific gravity of the hydraulic oil. Furthermore, the above formula 00 can be calculated by adding the hydraulic pressure correction value (margin value) ΔP1 to the hydraulic pressure P, 7° determined by the formula 0.
The line oil pressure PR, is determined, and its oil pressure correction value ΔP
I may be a constant value, but it may also be a function of vehicle operating state parameters. This oil pressure correction value ΔP1 is based on power loss and steady-state deviation Δe (corresponding to Δ■.), which are contradictory to each other.
is determined at the equilibrium point. That is, when ΔP1 is made small, the steady-state deviation becomes large, but when ΔP+ is made large, the steady-state deviation becomes small. However, ΔP.

The larger the pressure is, the more an unnecessarily large first line oil pressure PR will be generated under many operating conditions.

In the control method of the first line oil pressure pH and the second line oil pressure Plt in order to obtain the above target speed ratio e1 and maintain the tension of the transmission belt 24 at an appropriate level, the primary side oil pressure is The hydraulic pressure Pin in the cylinder 26 is the hydraulic pressure P in the secondary hydraulic cylinder 28. The tension of the transmission belt 24 is controlled exclusively by the second line hydraulic pressure Plt, which is made relatively high with respect to u'L.However, during engine braking, the hydraulic pressure P in the secondary side hydraulic cylinder 28 is increased.

1t is relatively high (relatively high) to the oil pressure P in the primary side hydraulic cylinder 26, and the tension of the transmission belt 24 is exclusively controlled by the first line oil pressure PR. That is, in step S7 If it is determined that the engine is braking, the direction of power transmission in the belt-type continuously variable transmission 14 is reversed, so that steps S10 and Sll, which are substantially similar to steps S8 and S9, are executed. Determine the oil pressure Pi° required in the primary hydraulic cylinder 26 and the first line oil pressure PR, and determine the oil pressure P0.° required in the secondary hydraulic cylinder 28.
And the second line oil pressure Pff2 for that purpose is determined.

Specifically, in step SIO, the optimum thrust force W i of the primary hydraulic cylinder 26 is determined based on the output torque T, and the speed ratio e from the pre-stored relationship shown in the following equation (2).
No is calculated, and the oil pressure Pi° to be supplied to the primary side hydraulic cylinder 26 is calculated from the following formula 〇3.
From the following formula 04), the above oil pressure p, fil and oil pressure correction value Δ
The first line oil pressure PR is calculated based on P2. This hydraulic pressure correction value ΔP is determined in order to accurately control the belt clamping force, and may be a constant value. Further, the oil pressure correction value ΔP2 in the above α-type is preferably calculated from a predetermined relationship. FIG. 8 shows the output pressure characteristics of the flow control valve 72 during engine braking. FIG. 8 shows the hydraulic pressure P1 in the primary hydraulic cylinder 26 and the hydraulic pressure P in the secondary hydraulic cylinder 28. This shows the change characteristics of ul with respect to the control value V0, and ΔV
, the thrust force is balanced and the speed ratio e is constant, then the oil pressure Pi, l in the primary side hydraulic cylinder 26 is the first
The value is smaller than the line oil pressure PL by ΔP2.

Therefore, the above-mentioned Δ
The first line oil pressure PE to be controlled by adding Pt
1 is required. This ΔP2 is the output oil pressure change characteristic of the flow rate control valve 72, and is the control value ■. , line oil pressure difference (PJ!I
Plt), but the control value v0 is (e”-e)/
e and the line oil pressure difference (P R+ −
Since P It ) is determined based on the output torque T and the speed ratio e, ΔP2 is a function of the speed ratio e, the target speed ratio e II, and the output torque T0, and these speed ratios e and the target speed ratio e* , output torque T, is obtained from the actual values.

wi,,' = f (T,,e)・
...Continued (In step Sll, the thrust ratio T- for obtaining the target speed ratio e1 is basically calculated based on the target speed ratio e1 and the output torque T from the image of the following formula α, and From the following formula 〇〇, the thrust force W.ut゛ of the secondary side hydraulic cylinder 28 to obtain the above thrust ratio γ- is obtained based on the thrust ratio γ- and the thrust force W!s° of the primary side hydraulic cylinder 26, and Secondary side hydraulic cylinder 28 from On type
The required oil pressure P 611L is found, and the above oil pressure P is obtained from the following formula 〇. ut.

The second line oil pressure PI is determined based on.

Note that since the second line oil passage 64 and the secondary hydraulic cylinder 28 are directly connected via the secondary oil passage 73, the hydraulic pressure correction value is zero.

7- = f (e”, T,) ・−−α
Wout'=γ--W i ,,'...
・αe In this way, the second line oil pressure Pl! , when the first line oil pressure pH is determined, step 3 is performed during forward drive.
12. During engine braking, step S13 is executed to calculate the deviation (e''-) between the target speed ratio e8 and the actual speed ratio e.
e) It is determined whether /e is positive or negative.

For example, in step S12 executed during normal drive, if it is determined to be positive, step S14 is skipped and step 315 is executed, whereby the first line oil pressure PR determined as described above and Second
Drive signals PDI, P to realize line oil pressure Pi,
The control amount v,,v,', which is the content of D2, is expressed by the following formula〇g
1 and C21, and then step 31
6 is executed, the signal is output to the first pressure regulating valve 68 and the second pressure regulating valve 70. In step S16, the control amount Vo determined in step S5 is also output to the flow rate control valve 72. However, the above step S1
If it is determined to be negative in step 2, step 31
4 is executed, the following equations (21) and (2)
Accordingly, the first line oil pressure PR becomes the second line oil pressure P12, and the second line oil pressure P12 becomes the first line oil pressure PR.
.

Steps S15 and S16 are executed after the exchange is performed. By repeating the above steps, the speed ratio e, the first line oil pressure PR, and the second line oil pressure P12 are controlled as shown in FIGS. 9 and 10.

V, = f (Pl+) ・ ・ ・α smell Vz
=fCpHt)...01Pj! +=
P1g ・ ・ ・+21) P lit
"P R+ . . . (2) Below, the operation of this embodiment will be considered in more detail.

First, when the drive is positive and the belt type continuously variable transmission 14 is in a steady state, Pin>Pout always holds, and the deviation (e"-e)/e is positive. This means that when a certain speed ratio is A certain cylinder oil pressure to maintain, e.g.
In order to obtain the cylinder oil pressure shown by the mark, the controlled variable Δv0 must be output at all times, but for this reason, the speed ratio always has a steady deviation, and the controlled variable V0
(=k・l e”−e l/e) becomes thicker (as the primary side cylinder oil pressure Pin increases, the speed ratio changes to the speed increasing side (increasing direction).For this reason, The above Δv0 corresponds to the speed ratio deviation formed on the speed increasing side compared to the current speed ratio, and in this case, the steady deviation (e”
-e)/e becomes positive. Therefore, in a steady state, step 312 is always determined to be positive, and steps SL5 and subsequent steps are directly executed.

In the same positive drive as described above, and in the speed ratio increasing change state, there is a large difference between the target speed ratio e1 and the actual speed ratio e, so the control amount V is determined by the above equation (4). increases, and the flow rate control valve 72 is opened. As a result, the seventh
As is clear from the figure, there is a margin pressure ΔP, so ■.

When P5 becomes larger, P5 increases with respect to P6ut, and an increasing speed change is executed. Figure 9 shows the changes in the speed ratio e, deviation (e''-e)/e, and oil pressure values of each part at this time.On the other hand, it is a forward drive, and the speed ratio is decelerated (decreased). Even in the state of change, since there is a large difference between the target speed ratio e8 and the actual speed ratio e, the control amount ■ becomes larger according to equation (4), and the flow rate control valve 72 is opened. Therefore, in order to supply hydraulic oil into the secondary hydraulic cylinder 28 and discharge hydraulic oil from the primary hydraulic cylinder 26, the value of the first line hydraulic pressure P11 determined in step S9 is and the value of the second line oil pressure P12 determined in step S8 are exchanged in step 314.In other words, since the deviation (e''-e)/e>Q is satisfied during the speed-up shift, the value of the second line oil pressure P12 determined in step S9 is exchanged. The value of the first line oil pressure Pl+ obtained in step S8 and the value of the second line oil pressure Pi obtained in step S8 are realized as they are, but the deviation (e''-e
)/e<Q, step 314 is executed, and the value of the first line oil pressure PR2 obtained in step S9 is used as the second line oil pressure, and the value of the second line oil pressure P12 obtained in step S8 is used as the second line oil pressure. The value is realized as the first line oil pressure. Therefore, step S14 performs the same function as the switching mechanism of a conventional speed change control valve. Figure 10 shows the speed ratio e and deviation (e
"-e)/e, which shows the change in the oil pressure value of each part. In FIG. As time passes, the actual speed ratio e approaches the target speed ratio e1, and eventually the deviation (el-e)/e becomes zero or slightly positive.At this time, the correlation between the line oil pressures returns to the original. After that, the speed ratio stabilizes at a constant value with a certain positive steady-state deviation, and the shift ends.

Further, even when the engine is braked, after the first line oil pressure pH and the second line oil pressure PR are determined in steps SIO and Sll, step S
13, the magnitude of the deviation (011-e)/e with respect to a predetermined judgment value cd is determined. FIG. 8 shows the output characteristics of the flow control valve 72 during engine braking. In a steady state where the speed ratio of the belt type continuously variable transmission 14 is maintained at a constant value, Δ
It is necessary to always output v0. That is, Δ■. corresponds to the steady-state error on the deceleration side. Step S1
3, if the deviation (e''-e)/e is less than or equal to the judgment value 04, it is considered to be a steady state or a deceleration shift state, and steps 315 and subsequent steps are executed as is, but if it is greater than the judgment value C4, the speed is increased. It is assumed that the deviation is in the state, and step S14 is executed. In the steady state, the deviation is a positive value of 06 or less, and in the deceleration shift state, the deviation is negative. In this case, the deviation is calculated in steps SIO and Sll. The first line oil pressure PR and the second line oil pressure P12 are realized as they are without being replaced.

In the steady state during engine braking of the belt type continuously variable transmission 14 or in the deceleration shift state, the first line oil pressure P1+
is always smaller than the second line oil pressure PR1, so the magnitude relationship between the first line oil pressure PII and the second line oil pressure Pi2 cannot be changed.

On the other hand, in the case of increasing speed, the primary hydraulic cylinder 26
It is necessary to make the thrust of
First line oil pressure P determined at IO and Sll
The magnitude relationship between R+ and the second line oil pressure PR1 is not appropriate, and it is necessary to switch the magnitude relationship in step 314. Step 313 is for this purpose.

Here, the characteristics of the flow control valve 72 shown in FIGS. 7 and 8 are determined by its structure. That is, since the spool valve element of the flow control valve 72 is assembled with a certain clearance, there is some leakage. Therefore, in the output port door 6, an intermediate pressure between the pH and atmospheric pressure that is applied to the input port door 5 is generated when the valve is closed, but the influence of the control pressure originally introduced as the flow control valve 72 is opened is strong. Eventually it will match.

Therefore, the first line oil pressure P11 is changed to the second line oil pressure P1.
The state shown in FIG. 7, which is greater than t, and the second line oil pressure P
The output characteristics of the flow rate control valve 72 are different from the state shown in FIG. 8 where R1 is larger than the first line oil pressure PR+. In this case, control accuracy will be higher if the low-pressure side cylinder oil pressure becomes the low-pressure side line oil pressure, so in this embodiment, the primary side oil path is set so that the characteristics shown in Fig. 7 are achieved during forward drive, which is frequently used. 71 is provided with a flow rate control valve 72.

As described above, according to this embodiment, the speed ratio e is basically controlled according to the first line oil pressure PR and the second line oil pressure P12, and in addition, the speed ratio e is controlled accurately by the flow rate control valve 72. is controlled by feedback. Furthermore, the flow rate control valve 7 can be controlled by the relatively simple configuration of the distribution valve 44.
Since there is no need to provide a switching function in the hydraulic control device 2, the configuration of the hydraulic control device becomes simple and inexpensive, and reliability is improved. For example, compared to the conventional hydraulic control device described in Japanese Patent Application No. 61-37571, in which the speed change control valve performs both flow rate control and pressure switching control, the spool valve is moved to the neutral position. Since there can be only one pair of springs to hold and one pair of solenoids to drive them, the speed change control valve is simple and inexpensive, and reliability is improved because the number of parts is reduced.

In addition, the above-mentioned distribution valve 44 allows the first
Since hydraulic oil can be supplied to the pressure regulating valve 68 and the second pressure regulating valve 70, the pressure regulating ranges of the first pressure regulating valve 68 and the second pressure regulating valve 70 may be the same, and valves with the same specifications may be used. Therefore, in this respect as well, the hydraulic control device becomes inexpensive and easy to manufacture.

Furthermore, even with the configuration of this embodiment, the advantages of the conventional hydraulic control device described above are not impaired. That is, the first line oil pressure P is adjusted by the first pressure regulating valve 68 and the second pressure regulating valve 70.
Since I and the second line hydraulic pressure PN are prepared, the pressure difference between them causes the primary side hydraulic cylinder 2 to
6 and the secondary side hydraulic cylinder 28, or the flow rate of the hydraulic oil discharged therefrom is determined.

Therefore, the speed ratio change speed is the belt type continuously variable transmission 14.
Regardless of the actual speed ratio and transmission torque (output torque T,), the first line oil pressure PR and the second line oil pressure Plt
Since it is determined according to the differential pressure between the two, sufficient transient response characteristics for speed ratio control can be obtained. Moreover, by controlling the first tFl pressure valve 68 in relation to the output torque T0 of the engine 10 and the actual speed ratio e, for example during normal drive, the first line oil pressure Pl can be adjusted to a sufficient speed ratio change speed. By controlling the second pressure regulating valve 70 in relation to the speed ratio and transmission torque, the second line hydraulic pressure P12 is controlled to a necessary and sufficient value so as to prevent power loss from occurring. Since the power loss is controlled to a necessary and sufficient value within a range that does not occur, there is an advantage that the power loss of the vehicle is significantly reduced.

Further, the flow rate control valve 72 includes the pair of pressure regulating valves 68 and 70.
1st line oil pressure P11 and 2nd line oil pressure P1
．． The hydraulic pressure regulated to the above-mentioned primary side hydraulic cylinder 26
Since it is provided in only one of the oil passages 71 and 71 of the pair of oil passages 71 and 73 that supply oil to the secondary side hydraulic cylinder 28, it is composed of a set of simple valves that can control the flow rate.
-The oil pressure control device is simple and inexpensive.

Although one embodiment of the present invention has been described above, the present invention is also applicable to other aspects.

For example, in the above-mentioned embodiment, the flow control valve 72 was provided only in the primary oil passage 71, which is one of the pair of primary oil passages 71 and secondary oil passages 73;
As shown in the figure, a flow control valve 72 is installed only in the secondary oil passage 73.
Even if it is provided, a certain effect can be obtained.

Furthermore, in the deceleration shift state during positive drive, which is determined to be negative in step S12, and the increase shift state during engine braking, which is determined to be greater than cd in step S13, both of the Although the value of the first line oil pressure Pl and the value of the second line oil pressure P12 are exchanged, these processes may be divided into different control formulas. That is, since there are cases where a rapid speed change is required for deceleration during forward drive, instead of simply exchanging the pH value and the PR2 value, PR4 and P1! Using equations (2) and (to), it is newly set to something that better meets the requirements. In addition, instead of that, the deviation (e
``-e)/e is smaller than a predetermined value C, the value of PR, and the value of PR2 are exchanged, but the deviation (e''
-e) If /e is larger than a predetermined value C, the value PR may be set as the lowest pressure and the value PR2 may be set as the maximum pressure.

P lx = Pj! ＋ ＋, □...(2) Also, there is no need to actively control the speed ratio during engine braking, and there is no problem even if the speed ratio is determined as it happens, so step S13 is omitted and step Sll is followed by Step S15 may also be executed. Even in this case, deceleration speed change is possible and sufficient practicality can be obtained.

Moreover, although the primary side hydraulic cylinder 26 and the secondary side hydraulic cylinder 28 of the belt type continuously variable transmission 14 described above have the same pressure receiving area, they may have different pressure receiving areas.

Further, in the above-described embodiment, the speed change control valve 44 was controlled so that the target speed ratio e1 and the actual speed ratio e matched, but the target rotational speed N in '' obtained in step S3 Rotational speed N i of the primary rotating shaft 16
n may be controlled so that they match, or it may be controlled so that the requested output of the vehicle and the actual driving force match.

In addition, in the above-mentioned example, as the deviation (e II-e)
Although the quantity /e was used, it may also be (e'' - e).

Further, in the above-described embodiment, the throttle valve opening degree was detected as the amount representing the required output of the vehicle, but in a vehicle equipped with a diesel engine or the like, an accelerator pedal operation amount or the like may be used.

Furthermore, in the above-described embodiment, the output of the flow control valve 72 was proportionally controlled by continuously changing the position of the spool valve 76. Controls may also be used.

Furthermore, if it is acceptable that the accuracy and stability of the speed ratio control decrease somewhat, the flow rate control valve 72 may be removed as shown in FIG. 12, that is, the first line oil pressure P1 ．． and the second line oil pressure Pl, as described above, are determined in order to obtain the target speed ratio e1 and set the tension of the transmission belt 24 to the optimum value, so the flow control valve 72 71 or the secondary oil passage 73, the speed ratio e and the tension of the transmission belt 24 can be controlled.

Further, the curve shown in FIG. 4 may be an optimum curve determined to achieve both fuel efficiency and drivability.

Note that the above-mentioned embodiment is merely one embodiment of the present invention, and various modifications may be made to the present invention without departing from the spirit thereof.

[Brief explanation of the drawing]

FIG. 1 is a diagram showing the configuration of an embodiment of the present invention. Second
This figure is a flowchart for explaining the operation of the embodiment shown in FIG. FIG. 3 is a diagram showing relationships used to explain the operation of the flowchart in FIG. 2. FIG. 4 is a diagram showing the minimum fuel consumption rate curve of the engine of FIG. 1. 5 and 6 are diagrams showing the change characteristics of the cylinder oil pressure with respect to the speed ratio in the embodiment shown in FIG. 1, respectively. FIG. 5 shows a positive torque state, and FIG. 6 shows an engine braking state. . FIGS. 7 and 8 are diagrams showing the output characteristics of the flow control valve shown in FIG. 1, respectively, with FIG. 7 showing the vehicle in a forward driving state, and FIG. 8 showing the engine braking state of the vehicle. There is. 9 and 10 are time charts respectively showing the transient characteristics of the oil pressure of each part in the positive torque state in the embodiment shown in FIG. Indicates the condition. FIGS. 11 and 12 are diagrams showing essential parts of other embodiments of the present invention. 14: Belt type continuously variable transmission 16: Primary rotating shaft 18: Secondary rotating shaft 20 Primary variable pulley 22: Secondary variable pulley 24: Transmission belt 26 Primary hydraulic cylinder 28: Secondary side Hydraulic cylinder 44: Distribution valve 68: First pressure regulating valve 70: Second pressure regulating valve 72: Flow rate control valve Applicant Toyota Motor Corporation Throntle fF Gate θth Engine rotation speed Ne Fig. 5 Fig. 6 ■-1! -1-! -L ■・km. 91st! iQ Figure 10

Claims (5)

[Claims] (1) 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 transmission belt that is wound around the pair of variable pulleys to transmit power; A hydraulic control device for a belt-type continuously variable transmission for a vehicle, comprising a pair of primary side hydraulic cylinders and a secondary side hydraulic cylinder that respectively change the effective diameters of the pair of variable pulleys, the device comprising: a pair of first output ports; and a first valve chamber and a second valve chamber for respectively discharging hydraulic oil to a second output port, the hydraulic pressure is adjusted such that a predetermined constant differential pressure is generated between the first valve chamber and the second valve chamber. a distribution valve that distributes hydraulic fluid pumped from a source and causes it to flow out from the first output port and the second output port; a pair of first line hydraulic pressures and a second line hydraulic pressure for obtaining a target speed ratio and maintaining an appropriate tension of the transmission belt; A hydraulic control device for a belt-type continuously variable transmission for a vehicle, comprising: a pressure regulating valve and a second pressure regulating valve. (2) One of the first pressure regulating valve and the second pressure regulating valve is configured to use a relatively lower hydraulic pressure of the first line hydraulic pressure and the second line hydraulic pressure for torque transmission of the belt type continuously variable transmission; The pressure is regulated to a value that generates sufficient tension in the transmission belt, and the other of the first pressure regulating valve and the second pressure regulating valve is set to a relatively higher side of the first line hydraulic pressure and the second line hydraulic pressure. The belt-type continuously variable transmission for a vehicle according to claim 1, wherein the hydraulic pressure is regulated to a value that generates a thrust necessary to maintain the speed ratio of the belt-type continuously variable transmission at the target speed ratio. Hydraulic control device. (3) One of the pair of pressure regulating valves supplies the working hydraulic pressure regulated to the first line hydraulic pressure or the second line hydraulic pressure to the primary hydraulic cylinder or secondary hydraulic cylinder via a flow control valve. A hydraulic control device for a belt-type continuously variable transmission for a vehicle according to claim 1 or 2. (4) The flow rate control valve is provided in an oil passage that supplies the first line hydraulic pressure regulated by the first pressure regulating valve to the primary hydraulic cylinder. Hydraulic control device for belt-type continuously variable transmissions for vehicles. (5) The flow rate control valve increases the flow cross-sectional area in accordance with the magnitude of the deviation between the actual speed ratio and the target speed ratio of the belt type continuously variable transmission, and Alternatively, the second pressure regulating valve increases the oil pressure of the one of the pair of primary variable pulley and secondary variable pulley having a relatively smaller thrust, or increases the hydraulic pressure of the primary variable pulley in response to the magnitude of the deviation. The hydraulic control device for a belt-type continuously variable transmission for a vehicle according to claim 3 or 4, which lowers the hydraulic pressure of the pulley and the secondary side variable pulley on the side having a relatively large thrust. .