JPS63173732A - Hydraulic control device of belt type continuously variable transmission for vehicle - Google Patents

Abstract

PURPOSE:To reduce a pump driving loss and improve fuel consumption by regulating first line hydraulic pressure in such a way as to generate said pressure causing a deviation between an actual speed ratio and a second target value to agree to the predetermined constant target deviation value. CONSTITUTION:A controller 94 comprises CPU 102, ROM 104 and RAM 106. The CPU 102 processes input signals according to a program preliminarily stored in the ROM 104, while using the memory function of RAM 106, and feeds a primary driving signal VD1 and a secondary driving signal VD2 respectively to a first pressure regulation valve 48 and a second pressure regulation valve 58 for controlling line hydraulic pressure Pl1 and second line hydraulic pressure Pl2. Concurrently therewith, speed ratio signals RA1 and RA2 for driving first and secondary electromagnetic solenoids 74 and 76 are fed thereto for controlling a speed ratio (e).

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.

As a hydraulic control system for such a vehicle belt-type continuously variable transmission, the applicant of the present application has previously filed Japanese Patent Application No. 61-3.
In No. 7571, +8) 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 (h
) By supplying the hydraulic oil whose pressure has been regulated to the first line hydraulic pressure to one of the primary hydraulic cylinder and the secondary hydraulic cylinder, and simultaneously causing the hydraulic oil in the other to flow out,
(C) a speed change control valve that adjusts the speed ratio of the continuously variable transmission by changing the effective diameters of the primary variable boob and the secondary variable boob; and (C) the primary hydraulic cylinder and and a second pressure regulating valve that regulates the pressure of the hydraulic oil flowing out from the other side of the secondary side hydraulic cylinder to a second line oil pressure lower than the first line oil pressure, and the valve has a second pressure regulating valve that adjusts the pressure of the hydraulic oil flowing out from the other side of the secondary side hydraulic cylinder to a second line oil pressure lower than the first line oil pressure. The present invention has proposed a type in which the speed change control valve is controlled so that the speed ratio is equal to the target speed ratio or the target input shaft rotation speed determined according to the driving state of the vehicle.

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, for example, the first line hydraulic pressure is adjusted to the target speed ratio during steady state. The drive side variable pulley thrust is generated to a necessary and sufficient value that does not cause loss of power, and also to a value that allows a sufficient speed ratio change speed, that is, a sufficient speed change response when changing gears. The first pressure regulating valve is controlled, while the second pressure regulating valve is controlled so that the second line oil pressure is at a necessary and sufficient value that does not cause slippage of the transmission belt, thereby minimizing the power loss of the vehicle. Sufficient transient response characteristics during gear shifting can be obtained while maintaining the speed at a low level.

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. 3, for example, and the oil pressures in both hydraulic cylinders P iR + P o
Assuming that the speed ratio ut is achieved at the oil pressure indicated by the mark, there is a steady state between that speed ratio and the target speed ratio that corresponds to ΔV0 (the amount of movement of the speed change control valve). A deviation occurs, but this steady deviation is
If the line oil pressure PR is increased, the slope of the hydraulic characteristics becomes steeper, so it becomes smaller, and if the first line oil pressure Pl is reduced, the slope of the oil pressure characteristics becomes gentler, so it becomes larger. However, if the first line oil pressure pH is increased (this will increase the driving loss of the pump, the excess oil pressure ΔP,
is determined at the balance point between driving loss and steady-state deviation, which are contradictory to each other.

On the other hand, due to individual differences in the parts that make up the belt-type continuously variable transmission, the predetermined thrust ratio characteristics and the actual characteristics of the continuously variable transmission do not necessarily match, and The thrust ratio characteristics of a 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. . Further, due to accuracy problems of the pressure regulating system including the first pressure regulating valve and the second pressure regulating valve, the first line oil pressure and the second line oil 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 to solve the above-mentioned problems, and its gist is that (a) the first pressure regulating valve; (C) a second pressure regulating valve, and the belt-type continuously variable transmission for vehicles is of the type that controls the speed change control valve so that the actual speed ratio matches the target value determined according to the driving state of the vehicle. A hydraulic control device for a machine, wherein the target value is corrected by a magnitude corresponding to a steady deviation of the feedback control so that the deviation becomes zero,
a correction means for creating a second target value as a direct target value of the feedback control; and adjusting a steady deviation between the actual speed ratio and the second target value to match a predetermined constant target deviation value. The control means includes a control means for feedback controlling the 111th pressure valve so as to regulate the first line oil pressure so as to generate the first line oil pressure.

In this way, the correction means corrects the target value by an amount corresponding to the steady-state deviation of the feedback control so that the deviation becomes zero, and the second target value is created. On the other hand, the first line hydraulic pressure is controlled by the control means to generate a first line hydraulic pressure that brings the steady deviation between the actual speed ratio and the second target value to a predetermined constant target deviation value. Oil pressure is regulated. Therefore, the actual deviation is made zero by the action of the correction means as described above, and the control means executes control to make the actual deviation from the second target value match the target deviation, so that the first line oil pressure is reduced. Controlled to a lower value. For this reason, compared to the conventional case where the first line oil pressure is controlled to the oil pressure obtained by adding the margin oil pressure in consideration of various errors, the driving loss of the pump and the power loss of the engine are reduced and the fuel efficiency of the vehicle is improved. can be improved.

Further, in the present invention, since the first line oil pressure is determined by the control means based on the steady deviation of the speed ratio, the pressure regulation accuracy of the first pressure regulating valve can be reduced, and the first tJR pressure valve can be manufactured at low cost. be able to.

The correction means preferably includes a storage means for storing a correction amount for reducing the deviation to zero in association with the driving condition of the vehicle, and uses the correction amount corresponding to the driving condition of the vehicle. The second target value is created by correcting the target value. It also has a storage means for storing a predetermined arithmetic expression, and calculates the second target value from the arithmetic expression based on the vehicle state. In addition, a correction value is calculated based on the driving condition of the vehicle from a predetermined arithmetic expression, but by learning and storing the correction amount at that time in relation to the driving condition, when the same driving condition occurs later, A second target value is immediately created according to the stored correction amount.

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 and a secondary rotating wheel 18, and a primary variable pulley 20 with a variable effective diameter attached to the primary rotating shaft 16 and the secondary rotating wheel 18. and a secondary variable pulley 22, a transmission belt 24 that is wound around the primary variable pulley 20 and the secondary variable pulley 22 to transmit power, and a primary variable boolean IJ 20 and the secondary variable pulley 22. It includes a primary hydraulic cylinder 26 and a secondary hydraulic cylinder 28 that change the effective diameter. The primary side hydraulic cylinder 26 and the secondary side hydraulic cylinder 28 are formed to have the same pressure receiving area, and the primary side variable pulley 20
The external shapes of the secondary variable pulley 22 are made the same, and the belt type continuously variable transmission 14 is made smaller. The primary variable pulley 20 and the secondary variable pulley 22 are
Fixed rotating bodies 31 and 32 fixed to the primary rotating ring 16 and the secondary rotating ring 18, respectively, are fixed to the primary rotating ring 16 and the secondary rotating ring 18, respectively, and are non-rotatable but movable in the axial direction. The fixed rotating body 31 is provided with
and movable rotating bodies 34 and 36 forming a V-groove between them.

The output from the secondary rotating shaft 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 48 . This oil pump 42
is driven by the engine IO via a drive shaft (not shown). The first pressure regulating valve 48 receives a first drive signal V, which will be described later.
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 according to DI, the first line oil pressure P! , to control.

This second line oil passage 52 is connected to the first discharge port 54 and the second discharge port 56 of the speed change control valve 44 and the second pressure regulating valve 5.
8, respectively. The second pressure regulating valve 5 is connected to the second line oil passage 5 according to the second drive signal VD2, which will be described later.
The second line oil pressure Pi, which is relatively lower than the first line oil pressure PR, is controlled by causing a part of the hydraulic oil in the second line to flow out to the drain oil path 60. 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 boat 54 and the second
Discharge boat 56. and oil passages 29 and 30 connected to the primary hydraulic cylinder 26 and secondary hydraulic cylinder 28.
a pair of first output ports 2 connected to each other via
and a cylinder bore 66 formed in the pulp body 65 so as to communicate with the second output capotro 4, one spool valve element 68 slidably fitted into the cylinder bore 66, and the spool valve. a pair of first springs 70 that hold the spool valve element 68 in the neutral position by biasing it toward the neutral position from both ends of the valve element 68;
and a second spring 72, and a first electromagnetic solenoid provided at both ends of the spool valve element 68 to continuously move the spool valve element 68 against the biasing force of the second spring 72 or the first spring 70. 74 and a second electromagnetic solenoid 76. On the 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 spool valve 68 is neutral as shown in the figure. When in position, it is formed at the same position as the first output capotro 2 and the second output capotro 4 in the axial direction of the spool valve element 68. Also, a position on the inner circumferential surface of the cylinder bore 66 that faces the pair of lands 80 and 82 when the spool valve element 68 is in the neutral position;
That is, at the positions where the first output capotro 2 and the second output capotro 4 open into the inner circumferential surface of the cylinder bore 66,
A pair of first annular grooves 86 and second annular grooves 88 with widths slightly larger than the lands 80 and 82 are formed.

The first annular groove 86 and the second annular groove 88 are connected to the land 80.
and 82 to form a constriction whose flow cross-sectional area changes continuously in order to control the flow of hydraulic oil.

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 boat 46 and the discharge boat 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 drain 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 boat 54 is continuously increased, the flow cross-sectional area between the second output capotro 4 and the input boat 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 thrusts 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 28, the primary hydraulic cylinder The hydraulic oil in 26 flows out, and the speed ratio e (rotational speed N of the secondary rotating shaft 18, ut/rotational speed N, 7 of the primary rotating shaft 16) of the belt type continuously variable transmission 14 becomes small.

Conversely, 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 boat 46 decreases. is continuously increased, while the flow cross-sectional area between the second output capotro 4 and the second discharge port 56 is increased. Hydraulic pressure is from the second output coupler.
The hydraulic pressure is higher than the working pressure output from the trolley 4 to the secondary hydraulic cylinder 28. For this reason, the belt type continuously variable transmission 1
4, the balance between the thrust forces of the primary hydraulic cylinder 26 and the secondary hydraulic cylinder 28 is disrupted, so while the hydraulic oil in the secondary hydraulic cylinder 28 flows out, the hydraulic oil flows into the primary hydraulic cylinder 26. The speed ratio e of the belt type continuously variable transmission 14 increases. In this way, the speed change control valve 4
4 has a switching valve function that supplies high-pressure hydraulic oil to one of the hydraulic cylinders 26 and 28 and low-pressure hydraulic oil to the other, and a flow control valve function that continuously adjusts the flow rate of the hydraulic oil. That's what I'm doing.

In this embodiment, a throttle oil passage 112 equipped with a throttle 110 for restricting the flow of hydraulic oil is connected between the second line oil passage 52 and the primary hydraulic cylinder 26 (connection oil passage 29). Furthermore, a throttle oil passage 116 equipped with a throttle 114 that restricts the flow of hydraulic oil is connected between the second line oil passage 52 and the secondary side hydraulic cylinder 28 (connection oil passage 30). .

As a result, the throttle oil passage 112 or the throttle oil passage 11
6, hydraulic oil is flowed from the high pressure side (drive side) hydraulic cylinder 26 or 28 to the second line oil passage 52,
As shown in FIG. 3, the output hydraulic pressure of the speed change control valve 44 (the hydraulic pressure Pt of the primary side hydraulic cylinder 26, and the hydraulic pressure P, ut of the secondary side hydraulic cylinder 28) is at the neutral position of the spool valve element 68. It is placed close to the 2-line hydraulic pressure P18. FIG. 4 shows such restricted oil passages 112 and 116.
4 shows the output hydraulic pressure characteristics of the speed change control valve 44 in a state where the shift control valve 44 is not provided. As is clear from comparison with Figure 4,
The hydraulic pressure curve Pin and the hydraulic pressure curve P out are generally lowered by the slight flow of hydraulic oil through the throttle oil passages 112 and 116 on the side to which hydraulic oil is not supplied from the first line oil passage 50 . Above aperture 1
10 and 114 have a necessary and sufficient flow area so that such an effect can be obtained and the loss of hydraulic oil will not be large. Therefore, the third
As shown in the figure, the hydraulic cylinder 26 on the low pressure side (driven side)
Alternatively, the oil pressure Pi or P out in 28 is made to substantially match the second line oil pressure Plz. Thereby, almost no differential pressure is generated between the second line oil pressure Pff and the oil pressure in the driven side hydraulic cylinder.

The belt-type continuously variable transmission 14 of the vehicle includes a primary rotating shaft 16.
a first rotation sensor 90 for detecting the rotation speed Ni of the
and a second rotation sensor 92 for detecting the rotational speed N0□ of the secondary rotating wheel 18.
From the rotation sensor 90 and the second rotation sensor 92, a rotation signal SRI representing the rotation speed N in and the rotation speed N are output.

A rotation signal SR2 representing 1L is output to the controller 94. The vehicle engine 10 also includes a throttle sensor 96 provided in its intake pipe for detecting the throttle valve opening θ, and an engine rotational speed N.

The throttle sensor 96 and the engine rotation sensor 98 send a throttle signal Sθ representing the throttle valve opening θ and a rotation signal SE representing the engine rotation speed N0 to the controller 94. Output to.

The controller 94 includes CPU 02, ROM 10
4. It is a so-called microcomputer including a RAM 106 and the like. The CPtJ102 processes the input signal according to a program stored in advance in the ROMIO4 while utilizing the memory function of the RAM106, and processes the input signal to control the first line oil pressure PR.
, and a first drive signal MDI to the first pressure regulating valve 48 and the second UR pressure valve 58 to control the second line oil pressure PR2.
and second drive signal VD2, respectively, and at the same time, speed ratio signals RAI and RA2 for driving the first electromagnetic solenoid 74 and the second electromagnetic solenoid 76 to control the speed ratio e are supplied to them.

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

First, by executing step S1, the rotation speed N1 of the primary rotation shaft 16, the rotation speed N0uL of the secondary rotation shaft 18, the throttle valve opening θ, and the engine rotation speed N are determined.
, are the rotation signals SRI and SR2, and the throttle signal Sθ
, are read into the RAM 106 based on the rotation signal SE. Next, in step S2, the target rotational speed N in '' of the primary rotation shaft 16 is calculated based on the throttle valve opening θ which is the required output of the vehicle from the relationship of the following equation (1) stored in advance in the ROM 104. This relationship is shown in FIG. 5, for example, and is determined in advance so that the engine 10 operates exclusively on the minimum fuel efficiency curve shown in FIG. Instead, a best fuel efficiency curve that is set to achieve both fuel efficiency and drivability may be used.

In step S3, the actual output torque T of the engine 10 is determined based on the actual engine rotational speed N and the throttle valve opening θth from the relationship of the following equation (2).

is calculated. Next, in step S4, the speed ratio e is determined from the following equation (3) by using the actual rotational speeds N i s and N. u
Calculated based on L. Then, in step S5,
From the following equation (4), the target speed ratio (
Target value) e* is calculated.

N52 = r (θLh) ・ ・
・ (1) T, = f(θi, N,) ・ ・
・(2) e = No□ /Ni ・ ・ ・(
3) e”=Nout/Nin”・
・ ・(4) In the following step S6, the speed change control valve 44
A second target speed ratio e° for controlling is calculated according to equation (5). Since the speed change control valve 44 always has a steady deviation due to its structure, in this step S6, the speed change control valve 44 is adjusted so that the speed ratio e achieved with this steady deviation becomes equal to the initial target speed ratio e1. The second target value e° is created by correcting the target value e1 for controlling 44.

However, C is a constant.

In step S7, a control value V0 for feedback controlling the speed change control valve 44 is determined based on the deviation between the second target speed ratio e° obtained in step S6 and the actual speed ratio C from the following equation (6). Calculated.

Vo −k (e' −e) /e −・・
(6) When the control value v0 is positive, the speed ratio signal R is set so that the spool valve element 68 is moved to the left and the rotational speed N out of the secondary rotating shaft 18 is increased.
A2 is output, and if it is negative, the spool valve 68
The speed ratio signal RAI is outputted so that the rotational speed N1fi of the primary rotating shaft 16 increases as the rotational speed N1fi of the primary rotating shaft 16 increases. Further, the magnitude of the speed ratio control value v0 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 element 68. Therefore, the above formula (
As is clear from 6), the speed ratio control value ■. is determined so as to match the actual speed ratio e and the second target speed ratio e°. Note that Equation (6) is a control constant. Therefore, the speed change control valve 44 adjusts the actual speed ratio e and the second target speed ratio e according to this speed ratio control value V0.
The feedback control is performed so that the second target speed ratio e.

is the direct target value of feedback control.

Here, the output oil pressure characteristics of the speed change control valve 44 are shown in FIG.
n > P out ) is realized, and if the oil pressure P in + P ouL at this time is the oil pressure indicated by the mark, respectively, then the speed ratio control value v
0 becomes Δv0, and this Δv0 is expressed by the above equation (6), so it is the second target speed ratio e.

This means that a steady deviation of a magnitude corresponding to Δv0 occurs between and the speed ratio e. As before, the target speed ratio e
When the speed ratio control value V0 is calculated based on the deviation between 8 and the speed ratio e, there is a difference of Δ between the target speed ratio e1 and the speed ratio e.
A deviation of a magnitude corresponding to v0 inevitably occurs, and in this case, the speed ratio e becomes a value smaller than the target speed ratio e1.

On the other hand, in this embodiment, an integral term (in this case, positive) is added to the target speed ratio e* in equation (5), resulting in a second target speed ratio e'' that is larger than the target speed ratio e1. is calculated, and the shift control value v0 is determined so that the second target speed ratio e° and the speed ratio e match, so the speed ratio e gradually increases and finally matches the target speed ratio e1. Note that the engine brake state (P
in<P out), the speed ratio e becomes larger than the target speed ratio e'', but since the integral term in equation (5) is negative, the speed ratio e gradually decreases and finally reaches the target speed ratio e''. The speed ratio is made to match the speed ratio e1.

Then, in step S8, it is determined whether the actual output torque T of the engine 10 is positive or not, that is, whether the engine 10 is in a positive torque state where power is being output or in an engine braking state. The reason why such a judgment is necessary is that the direction of power transmission is different between the positive torque state and the engine brake state, so the hydraulic cylinder 2
This is because the oil pressure change characteristics with respect to the speed ratio e of 6.28 change. For example, FIGS. 7 and 8 show the oil pressure P1 in the primary hydraulic cylinder 26 and the oil pressure P in the secondary hydraulic cylinder 28 in a positive torque state and an engine brake (negative torque) state. The hydraulic pressure change characteristics of ut are shown, and the magnitude relationship between the hydraulic pressure Pi and the hydraulic pressure P ouL 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 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, they appear directly in the magnitude relationship of hydraulic pressure.

If it is determined in step S8 that the output torque T is positive, step S9 is executed to generate a necessary and sufficient clamping force for the transmission belt 24 in the secondary hydraulic cylinder 28. The second pressure regulating valve 5 is operated so that the oil pressure (target oil pressure) Pout' can be obtained.
In step 8, the second line oil pressure PR to be regulated is determined.

That is, first, the following equation (
From the relationship 7), the actual output torque T of the engine lO,
Optimal secondary hydraulic cylinder 2 based on the actual speed ratio e
8 thrust (calculated value) Wo. Calculate 1°. Also, from the following equation (8), the above thrust force W. IIL' - pressure receiving area AQIIL of the secondary side hydraulic cylinder 28 + oil pressure (calculated value) based on the rotational speed N out of the secondary side rotating shaft 18 Po
ut.

At the same time, the above-mentioned oil pressure P is calculated from the following equation (9). uL
.

The second line oil pressure P12 is calculated based on.

W, uL f=f (T,,e)...(7
)...(8) P 1t =P,,t'...
- (9) Here, the above equation (7) 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 W
. ut' is the output torque T. and is varied in relation to the speed ratio e. Furthermore, in the relationship of equation (8), the second term is the rotation speed N. Oil pressure P is obtained by subtracting the centrifugal oil pressure that increases with ut from the first term. ut''. The second term 02 is the centrifugal force correction coefficient,
It is determined in advance from the specifications of the secondary side hydraulic cylinder 28 and the specific gravity of the hydraulic oil. In this embodiment, the second line oil passage 52 and the hydraulic cylinder 2 are connected through the throttle 110 or 114.
6 or 28, so the correction value can be changed to P.
. , t “.

On the other hand, if the output torque T of the engine 10 is O or negative in step S8, that is, if it is determined that the vehicle is in the engine braking state, the direction of power transmission in the belt type continuously variable transmission 14 is reversed. Therefore, in the negative torque state, the required oil pressure P in the primary side hydraulic cylinder 26
The second line oil pressure P12 is determined from "ea". That is, in step SlO, the following formula a stored in advance is determined.
From the relationship shown in [, the optimal thrust force Win° of the primary hydraulic cylinder 26 is calculated based on the output torque T, and the speed ratio e, and the hydraulic pressure p ia'' to be supplied to the primary hydraulic cylinder 26 is calculated from the following formula aυ. is calculated, while the second line oil pressure PI! is calculated from the following equation based on the above oil pressures p and fi+.
2 is calculated.

Wifi' = f (To, e)
--.001 When the second line oil pressure Pe is determined as described above, in step Sll, the pressure in the primary hydraulic cylinder 26 is adjusted to generate the necessary and sufficient thrust to achieve the target speed ratio. Oil pressure Pi or oil pressure P in the secondary hydraulic cylinder 28. The first line oil pressure PI to be regulated by the first pressure regulating valve 48 so that uL is obtained! , is determined according to the following equation α stored in advance in the ROM 104.

...α However, k, , T are constants. Further, ε is a target deviation (target value of deviation).

The first term on the right side of the above α temperature equation is a proportional operation order, and is a control amount to make the actual deviation le"-el/e with respect to the second target speed ratio e° match the target deviation ε, that is, the oil pressure value. Normally, such a proportional operation order increases the control amount relatively quickly, but it is inherently unable to eliminate steady control deviation.The second term on the right side of equation a1 is an integral operation term. A control amount, that is, a hydraulic pressure value, is generated to make the actual deviation 1e''-el/e from the second target speed ratio e° match the target deviation ε. Normally, such an integral action term increases the control amount relatively slowly, but it can reliably eliminate steady control deviations. Therefore, in this embodiment, the step Sll is a control means for controlling the first line oil pressure PR so that the deviation between the actual speed ratio e and the second target speed ratio e° matches the target deviation value ε. handle.

In this way, when the first line oil pressure pH is determined,
The next step 312 is executed and the following equations 0a and Q51
Accordingly, the first line oil pressure control value ■1 and the second line oil pressure control value v2 are determined. These formulas Oa and Q9
are the first line oil pressure PR determined in step S9 or SlO, and the second line oil pressure P
12 respectively, the first pressure regulating valve 48. The first line hydraulic pressure control value V1. This is to obtain the second line hydraulic pressure control value v2.

v, ==[(PβI)...'αaV
z = f (Pl!)...
・0 and the last step S in the series of steps
13, the speed ratio control value V6.13 determined in the previous step is determined. A first line oil pressure control value ■1 and a second line oil pressure control value ■2 are output.

As a result, as shown in FIGS. 7 and 8, the speed ratio e, the first line oil pressure PR, and the second line oil pressure Plt are controlled, and thereafter, steps S1 and subsequent steps are repeatedly executed.

As described above, in this embodiment, the integral value of the deviation between the target speed ratio e* and the actual speed ratio e is calculated as the target speed ratio e8.
According to equation (5), which is corrected by adding
° is calculated, and the speed ratio control value ■ of the speed change control valve 44 is set so that the second target speed ratio e° and the actual speed ratio e match. is determined. The correction term, which is the second term on the right side of (5) 2 above, increases the correction value in the direction of eliminating the deviation between the target speed ratio e* and the actual speed ratio e over time based on the size of the deviation. As a result, no steady deviation occurs between the second target speed ratio e° and the speed ratio e, and the speed ratio e is suitably controlled.

In addition, in this embodiment, since the first line oil pressure Pβ1 is determined according to the α temperature formula so that the second target speed ratio e° and the actual speed ratio e match in step Sll, the first line oil pressure PI can be controlled to a lower value than before, and the driving loss of the oil pump 42 and the running fuel efficiency of the vehicle can be further improved. That is, in steps S6 and S7, control is performed to make the target speed ratio e7 and the actual speed ratio e match, so that in step S11, for example, the deviation 1e - el/e is changed to a value (target deviation value). If feedback control is executed to make them match, there is a risk that both controls will interfere with each other, making it impossible to perform stable control. Further, since the deviation le"-el/e always occurs even if the target speed ratio e9 and the actual speed ratio e match, it can be treated as an independent quantity.

Therefore, if the first line oil pressure PR is controlled according to equation Q3 so that this deviation le'-el/e coincides with a certain target value, the first line oil pressure Pi.

is the value necessary to achieve the speed ratio that includes the target deviation ε, so it is lower than the conventional case where the first line oil pressure is calculated by adding the margin oil pressure to account for various errors. Controlled by value.

Moreover, in steps S6 and S7, the target speed ratio eI
Since the speed ratio deviation between ' and the actual speed ratio e can be eliminated, the target deviation value ε of 0 turbidity 2 can be set to a relatively large value, and in this respect as well, the first line oil pressure PR
+ can be set to a low value.

Next, the effects of this embodiment will be explained in comparison with a conventional hydraulic control device. FIG. 9 and FIG. 1θ show the prior art of this embodiment, and the patent application No. 61-37 filed earlier by the present applicant.
571 and Japanese Patent Application No. 61-172566, respectively. In the hydraulic control device of this embodiment, the actual deviation (e''-e)/e is eliminated by shifting the control target of the speed change control valve 44 from the original target speed ratio e8 to the second target speed ratio e°. Therefore, the target deviation of the first line oil pressure PJ and control can be increased.For this reason, as shown in FIG. 7, the first line oil pressure Pi is changed to the same as shown in FIGS. Here, the second target speed ratio e° determined in step S6 in the vicinity of the maximum speed ratio e + *ax and the minimum speed ratio emin is greater than 6+1111X, However, the actual speed ratio e cannot be larger than e,%,X or smaller than el,7.Due to such constraints, the deviation ( eo-el/e
The value of the first line oil pressure PR, which is necessary to make ε coincide with ε, becomes higher. In addition, in the conventional hydraulic control device having the characteristics shown in FIG. 10, the following formula OQ is provided in place of the speed ratio control formula in steps S6 and S7, and the following formula OQ is provided as the control formula for the first line oil pressure PN. η is provided.

In this case, the speed ratio steady deviation is the target deviation (maximum allowable value) without calculating the hydraulic pressure Pi in the high pressure side hydraulic cylinder.
The first line oil pressure P1. is determined, so
The first line oil pressure pH can be determined lower than in the conventional example shown in FIG. 9, but higher than in the embodiment shown in FIG. Further, a conventional hydraulic control device having the characteristics shown in FIG. 9 is constructed by removing the knot type from the conventional example shown in FIG. 1θ. In this case, the hydraulic pressure in the hydraulic cylinder on the high pressure side is the calculated value Pi
n'', a fairly large margin value ΔP1 is added to the calculated value P,.'' and the first line oil pressure Pl,
Determine. Therefore, the first line oil pressure Pl is high overall.

Vo =K (e” e) /e ・・・Ql・・
...αη In addition, in this embodiment, the steady-state deviation of the speed ratio 1e'-e
Since the first line oil pressure PR is determined by l/e, the pressure regulation accuracy of the first pressure regulating valve 48 can be reduced, and the first line oil pressure PR,
The pressure regulating valve 48 can be manufactured at low cost.

Although the embodiments of the present invention have 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, the speed ratio e and the target speed ratio e8 are configured to match, but the primary rotating wheel 16
rotational speed N1 and its target rotational speed (target value) N, ”
Exactly the same functions and effects can be achieved even if the configuration is made to match. In that case, the following equations θQl and (2111) are applied in place of the above equations (5), (6), and the tapping equation, respectively. In these equations, C and ' are proportional to each other. is a constant and a control constant, and N i n ” is a second target rotational speed (second target value).

...(181 V@=K' (N=, I Nin'') /Ni,
°...Q9)...(2) Also, in step s6 of the above embodiment, the second target value e° is calculated by the calculation of equation (5), but the deviation (
The correction amount S for making e''-e)/e zero is determined based on the driving condition of the vehicle, for example, the target speed ratio e1 and the required output torque T.
A large number of data maps are prepared in advance in connection with 1, and such data maps are stored in the ROM 104, and the second A target speed ratio may also be calculated. In this case, the correction amount S when the actual speed ratio e and the target speed ratio e11 match is calculated using the second term on the right side of equation (5), and this value is used to calculate the correction amount S that is already stored. You can update it.

This has the advantage that the correction amount can be updated to the optimum correction amount through learning.

Furthermore, in the equation (5) of the above embodiment, the second target speed ratio e°
Although only the integral term is used when calculating , it is also possible to add other terms to speed up the convergence.

Further, in the embodiment described above, even when the belt type continuously variable transmission 14 is in a speed change state, step S6 is executed and the correction term is added to the target speed ratio e8 to calculate the second target speed ratio e°. However, the purpose of the present invention is to reduce the deviation to zero in a steady state, and it is not necessarily necessary to correct the target speed ratio e1 in a shift state.

Further, in addition to the throttles 110 and 114 in the hydraulic circuit device of FIG. The discharge boat 54 may be a hydraulic circuit that is open to the atmosphere, and the speed change control valve 44 may be a hydraulic circuit that is open to the atmosphere.
It can also be constructed by a speed change control valve device composed of one or four spools.

Although other examples are not given, the present invention can be implemented with 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 device for a vehicle belt-type continuously variable transmission, which is an embodiment of the present invention. FIG. 2 is a flowchart for explaining the operation of the embodiment shown in FIG. FIG. 3 is a diagram showing the output oil pressure characteristics of the speed change control valve shown in FIG. 1. FIG. 4 is a diagram corresponding to FIG. 3 in the case where no throttle passage is provided between the second line oil passage and the hydraulic cylinder. 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. 7 and 8 are diagrams showing the change characteristics of the oil pressure of each part with respect to the speed ratio in the embodiment shown in FIG. 1, respectively. FIG. 7 shows the positive torque state, and FIG. 8 shows the engine braking state. There is. 9 and 10 are diagrams corresponding to FIG. 7 of a conventional hydraulic control device. 14: Belt type continuously variable transmission 16: Primary rotating shaft 18: Secondary rotating wheel 20 Primary variable pulley 22: Secondary variable pulley 24: Transmission belt 26 Primary hydraulic cylinder 28: Secondary side Hydraulic cylinder 44: Shift control valve 104: ROM (memory means) Step S6: Correction means Step S11: Control means Applicant Toyota Motor Corporation Fig. 4 (x+o”rρm) Fig. 5 Engine times? Figure 7 Figure 8 ta ratio e

Claims (4)

[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 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, a first pressure regulating valve that generates a first line hydraulic pressure; , the first
By supplying the hydraulic oil whose pressure has been regulated by the pressure regulating valve to one of the primary side hydraulic cylinder and the secondary side hydraulic cylinder, and simultaneously causing the hydraulic oil in the other side to flow out, the primary side variable pulley and the secondary side variable pulley are a speed change control valve that adjusts the speed ratio of the belt type continuously variable transmission by changing the effective diameter of the transmission; and pressure of hydraulic fluid flowing out from the other of the primary hydraulic cylinder and the secondary hydraulic cylinder through the speed change control valve. and a second pressure regulating valve that adjusts the pressure to a second line oil pressure lower than the first line oil pressure, the target value determined based on the operating state of the vehicle and the actual speed change of the belt type continuously variable transmission. A hydraulic control device of a type that performs feedback control of the speed change control valve so that a deviation from a ratio or an actual rotational speed of the primary rotating shaft becomes small, and the feedback control is performed in a steady state so that the deviation becomes zero. a correction means for creating a second target value as a direct target value of the feedback control by correcting the target value by a magnitude corresponding to the deviation; a control means for controlling the first pressure regulating valve to generate a first line oil pressure for making a steady deviation from a target value coincide with a predetermined target deviation; Hydraulic control device for continuously variable transmission. (2) When the speed ratio is e, the target value is e^*, and the deviation thereof is (e^*-e)/e, the correction means
The second target value e° is determined according to the following formula ▲ There are mathematical formulas, chemical formulas, tables, etc. ▼ (where C is a proportionality constant). Transmission hydraulic control device. (3) The correction means has a storage means for storing a correction amount for reducing the deviation to zero in association with the driving condition of the vehicle, and uses the correction amount corresponding to the driving condition of the vehicle. The hydraulic control device for a belt-type continuously variable transmission for a vehicle according to claim 1, wherein the second target value is created by correcting the target value. (4) The correction means determines the rotational speed of the primary rotation shaft as N_i_n, the target value as N_i_n^*, and the deviation thereof as (N_i_n-N_i_n^*)/N_i_n^.
*, then the second target value N_i_n is determined according to the following formula ▲ There are mathematical formulas, chemical formulas, tables, etc. ▼ (where C is a constant of proportionality). A hydraulic control device for the vehicle belt-type continuously variable transmission described above.