JPS63279937A - Oil pressure control device for belt type non-stage transmission of car - Google Patents

Abstract

PURPOSE:To prevent increase in a driving loss of a pump due to the excess pressure rise of line oil pressure by providing a correcting means for actuating signal to correct the actuating signal to a value smaller than its actual value when the actuating signal becomes larger than a predetermined definite value. CONSTITUTION:Hydraulic oil returning through a return passage to an oil tank 38 is drawn through a strainer 40 and an oil drawing passage 41 into an oil pump 41, and then forcedly fed into the primary line oil passage 50 connected to an input port 46 of a gear shifting control valve 44 and the primary pressure regulating valve 48. The oil pump 42 is driven through a driving shaft by an engine 10, and the primary pressure regulating valve 48 causes the hydraulic oil in the primary line oil passage 50 to flow out partially into the secondary line oil passage 52 according to the primary driving signal VD1 to regulate oil pressure in the primary line oil passage 50 to the primary line oil pressure Pl1.

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 a vehicle, and in particular, it controls line hydraulic pressure based on a speed ratio deviation between an actual speed ratio and a target speed ratio. The present invention relates to an improvement of a hydraulic control device as described above.

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 device for such a belt-type continuously variable transmission for a vehicle, the hydraulic fluid is controlled to flow in and out of at least one of the primary hydraulic cylinder and the secondary hydraulic cylinder, thereby controlling the hydraulic pressure of the continuously variable transmission. and a pressure regulating valve that regulates line oil pressure supplied to the shift control valve, and determines the actual speed ratio of the continuously variable transmission according to the driving condition of the vehicle. A system has been proposed in which the speed change control valve is controlled so as to deviate from the set target speed ratio. For example, JP-A-61-218862
Publication No. 61-375 filed earlier by the applicant
The device described in No. 71, Japanese Patent Application No. 61-37576, etc. is an example of such a system. By controlling the pressure regulating valve so that the pulley thrust is generated to a necessary and sufficient value without causing power loss, and so that a sufficient speed ratio change speed, that is, a sufficient speed change response is obtained during gear changes, The advantage is that sufficient transient response characteristics during gear shifting can be obtained while maintaining the power loss as low as possible.

Incidentally, in such a hydraulic control device, it is inevitable that a deviation occurs between the actual speed ratio and the target speed ratio due to the output hydraulic characteristics of the speed change control valve, but this speed ratio deviation is always greater than a certain value. If you try to control it so that it is smaller, it is necessary to set the line oil pressure to a large value in advance, taking into account individual differences in continuously variable transmissions, changes over time, pressure adjustment errors of pressure regulating valves, etc. However, the reduction in power loss has not always been completely satisfactory. For this reason, the applicant further proposed in Japanese Patent Application No. 172566/1983 that a line hydraulic We have proposed a hydraulic control device in which the pressure regulating valve is feedback-controlled using the difference between the speed ratio deviation and the target deviation value as an operation signal so as to obtain the following. In this way, the line oil pressure will be set to the minimum level necessary to achieve a speed ratio that includes the target deviation value, regardless of individual differences in continuously variable transmissions, changes over time, pressure adjustment errors of pressure regulating valves, etc. Since it is hydraulically controlled, the power loss of the engine is reduced and the fuel efficiency of the vehicle is improved, and the accuracy required for the pressure regulating valve, speed change control valve, etc. is relaxed.

Problems to be Solved by the Invention However, the feedback control of such a pressure regulating valve is
This is to match the speed ratio deviation in a steady state where the speed ratio is approximately constant with the target deviation value. Although the ideal line oil pressure can be obtained in steady state, the actual speed ratio deviation and the target deviation value are obtained during shifting when the speed ratio changes. The difference with the deviation value, that is, the operating signal, becomes much larger than in the steady state, so the line oil pressure becomes higher than necessary, which increases the drive loss of the pump and causes sudden changes in the speed ratio. There was a problem that the driving maneuverability was impaired. For this reason, feedback control of the pressure regulating valve has conventionally been performed only during steady state conditions, but this is not always satisfactory because the control type changes between steady state and shift state. Although the degree of the difference becomes smaller, it still becomes higher than necessary near the boundary between the steady state and the shifting state.

Means for Solving the Problems The present invention has been made against the background of the above circumstances.
The purpose of this is to enable appropriate line oil pressure to be obtained through feedback control of the pressure regulating valve even during gear changes.

In order to achieve this object, the present invention controls the speed change control valve so that the actual speed ratio matches the target speed ratio determined according to the driving condition of the vehicle, as described above, and In order to obtain the line oil pressure to match the speed ratio deviation between the actual speed ratio and the target speed ratio with a predetermined constant target deviation value, the difference between the speed ratio deviation and the target deviation value is used as an operating signal. In a hydraulic control device for a vehicle belt-type continuously variable transmission that performs feedback control of a pressure regulating valve, when the operating signal becomes larger than a predetermined constant value, the operating signal is set to a value smaller than the actual value. The present invention is characterized in that an operation signal modification means is provided for modifying the motion signal.

Operation and Effects of the Invention In such a hydraulic control device, when the operating signal is smaller than a predetermined constant value, that is, when the speed ratio of the continuously variable transmission is steady and approximately constant, the pressure regulating valve is activated as in the conventional case. Feedback control is performed according to the operating signal, but if the operating signal becomes larger than a predetermined constant value, that is, the speed ratio of the continuously variable transmission changes, the speed ratio deviation increases and the target deviation value When the difference increases, the operating signal is modified by the operating signal modifying means to a value smaller than the actual value, and the pressure regulating valve is feedback-controlled in accordance with the modified value. For this reason, the pressure regulating valve is feedback-controlled to ensure that the appropriate line oil pressure is obtained not only when the continuously variable transmission is in steady state but also when changing gears, which increases the drive loss of the pump due to excessive increase in line oil pressure. This prevents deterioration in driving performance and maneuverability.

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 wheel 16 of a belt type continuously variable transmission 14 via a clutch 12. The belt type continuously variable transmission 14 includes a primary rotating shaft 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 shaft 18. and the secondary variable pulley 22, the transmission belt 24 that is wound around the primary variable pulley 20 and the secondary variable boolean IJ22 to transmit power, and the effectiveness of the primary variable boot 920 and the secondary variable pulley 22. It includes a primary hydraulic cylinder 26 and a secondary hydraulic cylinder 28 whose diameters are changed.

These primary side hydraulic cylinder 26 and secondary side hydraulic cylinder 28 are formed to have the same pressure receiving area,
The above primary side variable boolean I720 and secondary side variable pulley 2
The belt type continuously variable transmission 14 is made smaller because the outer diameters of the belts 2 and 2 are the same. The primary variable pulley 20 and the secondary variable pulley 22 are connected to fixed rotating bodies 31 and 32 fixed to the primary rotating wheel 16 and the secondary rotating shaft 18, respectively, and the primary rotating shaft 16 and the secondary rotating shaft 18, respectively. The movable rotary bodies 34 and 36 are provided on the next rotating ring 18 so as to be non-rotatable but movable in the axial direction, and form grooves between them and the fixed rotary bodies 31 and 32. In addition,
The output from the secondary rotating wheel 18 of the belt-type continuously variable transmission 14 is transmitted to the drive wheels of the vehicle via an auxiliary transmission, a differential gear device, etc. (not shown).

A hydraulic control circuit for operating the belt type continuously variable transmission 14 configured as described above is configured as described below. That is, the hydraulic oil that has returned to the oil tank 38 via a return path (not shown) is sucked into the oil pump 42 via the strainer 40 and the suction oil path 41, and is connected to the input boat 46 of the speed change control valve 44 and the first pressure regulating valve 48. The oil is fed under pressure to the first line oil passage 50. oil pump 42
is adapted to be driven by the engine 10 via a drive shaft (not shown), and the first pressure regulating valve 48 is driven by the first line oil passage 50 in accordance with a first drive signal VDI, which will be described later.
By causing a part of the hydraulic oil in the line to flow out to the second line oil passage 52, the oil pressure in the first line oil passage 50 is regulated to the first line oil pressure PR. In this embodiment, this first pressure regulating valve 4
8 corresponds to a pressure regulating valve, and the first line oil pressure PR corresponds to the line oil pressure. Further, the second line oil passage 52 is connected to the first discharge boat 54 and the second discharge boat 5 of the speed change control valve 44.
6 and a second pressure regulating valve 58, respectively. Second
The 11-pressure valve 58 directs a portion of the hydraulic oil in the second line oil passage 52 to the drain oil passage 6 in accordance with a second drive signal VD2, which will be described later.
0, the oil pressure in the second line oil passage 52 is regulated to a second line oil pressure Pl that is relatively lower than the first line oil pressure Pi. The first pressure regulating valve 48
The second #A pressure valve 58 is composed of a so-called electromagnetic proportional relief valve.

The speed change control valve 44 is a so-called proportional control solenoid valve, and the manual port 46. The first discharge boat 54 and the second
Discharge boat 56. so as to communicate with a pair of first output capotros 2 and second output capotros 4, which are connected to the primary side hydraulic cylinder 26 and the secondary side hydraulic cylinder 28 via connection oil passages 29 and 30, respectively. A cylinder bore 66 formed in a valve body 65, a spool valve element 68 slidably fitted into the cylinder bore 66, and a spool valve element 68 that is biased toward a neutral position from both ends of the child 68. A pair of first springs 70 and a second spring 72 are provided at both ends of the spool valve 68 to hold the spool valve 68 in a neutral position. The electromagnetic solenoid 74 includes a first electromagnetic solenoid 74 and a second electromagnetic solenoid 76 that are moved against the biasing force of the first spring 70. 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 when the spool valve 68 is in the neutral position, The first output capotro 2 in the axial direction of the valve element 6B
and is formed at the same position as the second output port 4. 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, the first output capotro 2
A pair of first annular grooves 86 and a second annular groove with a width slightly larger than the lands 80 and 82 are located at the position where the second output captro 4 opens into the inner circumferential surface of the cylinder bore 66.
An annular groove 88 is formed. The first annular groove 86 and the second annular groove 88 form a constriction whose flow cross-sectional area changes continuously in order to control the flow of hydraulic oil between the lands 80 and 82.

As a result, when the spool valve 68 is in the neutral position, the first output capotro 2 and the second output capotro 4
are evenly communicated with the input port 46 and the discharge bow (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 fluid is allowed to exit from the exhaust port 54,56.

However, as the spool valve element 68 is moved from the neutral position in one axial direction, for example, in the direction approaching the second electromagnetic solenoid 76, that is, in the right direction in the figure, the first output capotro 2 and the first discharge boat 54 is continuously increased, while the flow cross section between the second output capotro 4 and the input port 46 is continuously increased. The hydraulic pressure output to the hydraulic cylinder 26 is lower than the hydraulic pressure output from the second output capotro 4 to the secondary 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 wheel 18, ut/rotational speed N!7 of the primary rotating shaft 16) of the belt type continuously variable transmission 14 becomes small.

On the other hand, as the spool valve element 68 is moved from the neutral position toward the first electromagnetic solenoid 74, that is, to the left in the figure, the flow cross-sectional area between the first output capotro 2 and the input port 46 decreases. is continuously increased, while the flow cross-sectional area between the second output capotro 4 and the second discharge boat 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 flows into the primary hydraulic cylinder 26, the hydraulic oil in the secondary hydraulic cylinder 28 flows out. The speed ratio e of the belt type continuously variable transmission 14 increases. The speed change control valve 44 of this embodiment is as follows:
In this way, it 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. are doing.

On the other hand, between the second line oil passage 52 and the connection oil passage 29°30, there are throttles 110 for restricting the flow of hydraulic oil.
．． Restricted oil channels 112 and 116 with 114 are connected. As a result, those throttle oil passages 112 or 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,
The output oil pressure of the speed change control valve 44, that is, the oil pressure P in the primary side hydraulic cylinder 26! , , and secondary hydraulic cylinder 28
As shown in FIG. 2, the internal oil pressure Pout approaches the second line oil pressure P1t in condition B (speed ratio control value V(1=o) in which the spool valve 68 is held at the neutral position. and the hydraulic pressure P1 or P0 in the low pressure side (driven side) hydraulic cylinder 26 or 28.

is made to substantially match the second line oil pressure PIt.

The belt-type continuously variable transmission 14 of the vehicle also includes a first rotation sensor 90 for detecting the rotation speed N in of the primary rotation shaft 16, and a first rotation sensor 90 for detecting the rotation speed N, □ of the secondary rotation wheel 18. A second rotation sensor 92 is provided for
The first rotation sensor 90 and the second rotation sensor 92 output a rotation signal SRI representing the rotation speed N and a rotation speed N. , , t is output to the controller 94 . Further, the engine 10 is provided with a throttle sensor 96 for detecting the throttle valve opening θ as a quantity representing the required output of the vehicle, and an engine rotation sensor 98 for detecting the engine rotation speed N1.
The throttle sensor 96 and the engine rotation sensor 98 output a throttle signal Sθ representing the throttle valve opening θ and a rotation signal SE representing the engine rotational speed N to the controller 94.

The controller 94 includes the CPU 102. ROM104
．． The CPU 102 is a so-called microcomputer including a RAM 106, etc., and processes input signals according to a program stored in advance in the ROM 104 while utilizing the storage function of the RAM 1. speed ratio signals RAI and RA for driving the first electromagnetic solenoid 74 and the second electromagnetic solenoid 76 to control the speed ratio e while supplying the first drive signal VDI and the second drive signal VD2 respectively to
2 to them.

Hereinafter, the operation of this embodiment will be explained according to the flowchart shown in FIG.

First, by executing step S1, the rotational speed N i R of the primary rotating wheel 16 - the rotational speed N0ut* of the secondary rotating shaft 1B, the throttle valve opening θLh+, and the engine rotational speed N0 are determined by the rotational signals SRI and SR2. , throttle signal Sθ twice based on light signal SE
6 is read in. Next, in step S2, the RO
Based on the well-known relationship stored in M104, the actual output torque T1 of the engine 10 is determined based on the throttle valve opening θ and the engine speed N, and further in step S3, the throttle valve opening θ is determined. Based on this, the target rotational speed N i h '' of the primary rotation shaft 16 is determined.The relationship for determining the target rotational speed N2 is, for example, as shown in FIG. 4, and as shown in FIG. This is determined in advance so that the engine 10 operates exclusively on the minimum fuel efficiency curve.

Continuing (in step S4, the rotational speeds N i n and N. u
The actual speed ratio e of the continuously variable transmission 14 is calculated from t, and in step S5, the target speed ratio e3 is calculated from the rotational speed N o u & and the target rotational speed NLJ.

Then, in step S6, a speed ratio control value V0 for operating the speed change control well 44 so that the speed ratio e matches the target speed ratio e1 is set to a speed ratio control value V0 of the speed ratio e and the target speed ratio e'''. and the following equation (1) based on 0 for the control constant
Calculated according to In step S16, which will be described later, this speed ratio control value ■.

If is positive, the spool valve 68 is moved to the left and the rotational speed of the secondary rotating wheel 18 reaches N0.

The speed ratio signal RA2 is outputted so that the speed ratio signal RA2 increases, and when the speed ratio signal RA2 is negative, the spool valve element 68 is moved to the right and the speed ratio signal RA2 is outputted so that the rotational speed Ni of the primary side rotating shaft 16 increases. RAI is output. Also, the magnitude of the speed ratio control value v0 is determined by the speed ratio signal RAI or the speed ratio signal R.
This corresponds to the size of A2, that is, the amount of movement of the spool valve 68.

Vo = Ko (e” e) /e ・・
(tl) Then, in step S7, an operating signal is calculated according to the following equation (2) from the speed ratio deviation le"-el/e between the actual speed ratio e and the target speed ratio e1 and the predetermined target deviation value ε. Z is calculated. This operation signal Z is originally applied as is to the control equation 1161 in step 314, which will be described later, but in this embodiment, it is further applied to the following equation (3) stored in advance in the ROM 104 in step S8. ) is used.The relationship of equation (3) is as shown in FIG. 6, for example, when the operation signal 2 is When the actual operating signal 2 is smaller than the fixed value Zl, the actual operating signal 2 is used as the corrected value Z, and when the operating signal 2 exceeds the fixed value Z1, the value is smaller than the actual operating signal Z shown by the dashed line in the figure. The smaller value is taken as the correction value Z, and when the operation signal Z exceeds a predetermined constant value 21 which is larger than the above-mentioned constant value ZI, the correction value Z is set to 0.The operation signal Z is set to the constant value ZI.
If the speed ratio deviation le''-el/e is small and the difference from the target deviation value C is small, the continuously variable transmission 14
This is a steady state in which the speed ratio e of is approximately constant, and the operating signal Z
is larger than the constant value Z1, the speed ratio deviation le”−
This is a quasi-steady state in which el/e is larger than the steady state, and if the operating signal Z is larger than the constant value Zz, the speed ratio deviation +6"-el/e is in a shifting state which is significantly large. In the quasi-steady state, that is, when the operating signal Z is between the constant value 2 and Z2, the amount of modification of the operating signal Z becomes continuously large, and the modified value Z is It is continuously changed until it reaches 0.

ZC=f (Z)...
(3) Step S9 is then executed, and it is determined whether the actual output torque T8 of the engine 10 is positive or not, that is, whether it is in a positive torque state where power is output from the engine IO or in an engine braking state. be done. The reason why such a determination is necessary is because the power transmission direction is different between the positive torque state and the engine brake state, so the oil pressure change characteristic with respect to the speed ratio e of the hydraulic cylinders 26, 28 changes. If it is determined that the output torque T is positive, first, step S10 is executed, whereby the secondary hydraulic cylinder 28 is used to generate the necessary and sufficient clamping force on the transmission belt 24. The second line oil pressure P1z to be regulated by the second pressure regulating valve 58 is determined so that the oil pressure within (target oil pressure) Pout° is obtained. That is, first, the actual output torque T of the engine 10 is determined from the relationship of the following equation (4) stored in advance in the ROM 104, and the optimal thrust of the secondary side hydraulic cylinder 28 (calculated value) based on the actual speed ratio e. Wol° is calculated, and this thrust force W is calculated from the following equation (5). The second line oil pressure Pet is calculated based on the pressure receiving area A0□ of the secondary side hydraulic cylinder 28, and the rotational speed N of the secondary side rotating shaft 18 uL.The oil pressure determined by this equation (5) teeth,
Originally, the target oil pressure P0. However, in this embodiment, since the second line oil passage 52 and the hydraulic cylinder 28 are connected via the throttle oil passage 116,
The oil pressure expressed by equation (5) may be used as the second line oil pressure Pit.

Wout' = f (Ts, e)
H1+ (4)A(ut) Here, the above equation (4) is calculated in advance in order to set the tension of the transmission belt 24, that is, the clamping force on the transmission belt 24, to a necessary and sufficient value, Thrust W.ut
° is varied in relation to the output torque T and the speed ratio e. In addition, in the relationship of equation (5), the second term is calculated by subtracting the centrifugal oil pressure that increases with the rotational speed N ovt from the first term to obtain the second line oil pressure PR□ (P,, t
”).

The second term C7 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 S11, the first pressure regulating valve 48 is adjusted so as to obtain the oil pressure (target oil pressure) Pi in the primary side hydraulic cylinder 26 to generate necessary and sufficient thrust to realize the target speed ratio. A pre-programmed value Pi,... of the first line oil pressure to be regulated is determined by the following equation (6) stored in advance in the ROM 104.
), the actual output torque T of the engine 10,
and the thrust ratio γ during forward drive based on the target speed ratio e1,
(The thrust force W of the secondary hydraulic cylinder 28, ut/the thrust force W!M of the primary hydraulic cylinder 26) is calculated, and
From the following equation (7), the thrust ratio γ and the thrust W0 of the secondary hydraulic cylinder 28 are determined. The thrust force Win' of the primary hydraulic cylinder 26 is determined from the angle. Then, from the following equation (8), the thrust force W of the primary side hydraulic cylinder 26! +l'l Pressure receiving area A of the primary side hydraulic cylinder 26, 7. Rotational speed N of the primary rotating shaft 16! The oil pressure (calculated value) Pll is calculated based on fi, and the first line oil pressure PR11P) is calculated from the following equation (9) based on the oil pressure PiR' and the margin oil pressure ΔP.

T+ = f (T*, e”,) ・
...(6) γ.

Equation (6) above calculates the thrust ratio T over a wide range of operating conditions. This shows a relationship determined in advance so that the target speed ratio e'' and the actual output torque T can be determined in relation to the thrust ratio T. This is to find the line oil pressure P11 (P).In addition, in the relationship of equation (8) above, the second term is corrected by subtracting the centrifugal oil pressure, which increases with the rotational speed N!fi, from the first term. 01 in the second term is determined in advance from the specifications of the primary side hydraulic cylinder 26 and the specific gravity of the hydraulic oil.Furthermore, the above equation (9) calculates that the oil pressure Pi° obtained by the equation (8) has a margin oil pressure. The first line oil pressure P I I (P) is determined by adding ΔP.

Here, the above-mentioned surplus oil pressure ΔP is the speed ratio e and the target speed ratio e8
This is necessary in order to reduce the speed ratio deviation le'-e l/e. That is, the output oil pressure characteristics of this embodiment are shown in FIG. 2, and the speed ratio control value v
0 is expressed by the above equation (11), so when the speed ratio control value v6 is 0, the actual speed ratio e and the target speed ratio e*
However, in other cases, there is a speed ratio control value ■ between the actual speed ratio e and the target speed ratio e8. Speed ratio deviation 1e*- of the magnitude corresponding to
This results in el/e. This speed ratio deviation l e*-
When the first line oil pressure PR is increased, el/e becomes smaller because the slope of the oil pressure characteristics becomes steeper, and when the first line oil pressure Pi is made smaller, the slope of the oil pressure characteristics becomes gentler, so el/e becomes larger. However, as the first line oil pressure PII increases, the drive loss of the pump 42 also increases accordingly.
The margin oil pressure ΔP is determined at a balance point between the drive loss and the steady-state deviation, which are contradictory to each other.

In addition, the oil pressure P obtained from the above equations (6) 2 to (8)
i7° may vary due to individual differences in the parts that make up the continuously variable transmission 14, changes over time, pressure regulation errors in the first pressure regulating valve 48, etc.
Primary side hydraulic cylinder 2 at target speed ratio e8
It does not completely match the actual oil pressure P1 of 6, but the actual oil pressure P! When fi is smaller than P jn , the deviation 1e'-el/e becomes larger. Therefore, when controlling the 181st pressure valve 48 so as to obtain the first line oil pressure Pβ, , P), the margin oil pressure ΔP is determined by the error or adjustment of the oil pressure (calculated values) P, , ′ mentioned above. It was necessary to set the excess oil pressure in advance in anticipation of pressure adjustment errors in the pressure system, but in this embodiment, a feedback ring is added in step 314, which will be described later, so it is not necessary to take these errors into consideration when setting the margin oil pressure. There is no need to set ΔP thick.

On the other hand, if it is determined in step S9 that the vehicle is in the engine braking state, the direction of power transmission in the belt-type continuously variable transmission 14 is reversed, so steps S12 and 311, which are substantially the same as steps SIO and 311, are performed. By executing S13, the second line oil pressure P1
2 and 1st line oil pressure preprogram value P R+
Determine (F). That is, in step S12, the optimum primary side hydraulic cylinder 2 is determined based on the output torque Tel speed ratio e from the relationship shown in the following equation Ql stored in advance.
The thrust force Wifl' of 6 is calculated, and the hydraulic pressure pifi1 to be supplied to the primary side hydraulic cylinder 26 is calculated from the following equation part.
That is, the second line oil pressure PR is calculated.

In addition, in step S13, output torque T0. Calculate the thrust ratio T- based on the target speed ratio e, and calculate the thrust force of the secondary hydraulic cylinder 28 to obtain the thrust ratio γ- from the following equation 〇31.
' is determined based on the thrust ratio γ- and the thrust force Win' of the primary hydraulic cylinder 26, and the required oil pressure P in the secondary hydraulic cylinder 28 is determined from equation Q4. ut.

In addition, from the following formula Q51, the above hydraulic pressure P0, '
and the first line oil pressure P111 based on the margin oil pressure ΔP.
Calculate PI.

Wi, 11=f (T1.e) ...0ω
γ−=f (T,,e”) ...(2
) Wout'=w! , 11・r−・−・T13...a
Shun P 11 (?) = P out ′+ΔP
...-QSI In this way, the second line oil pressure P1
2 and 1st line oil pressure P j! When I (P) is calculated, the next step 314 is executed, and a feedback ring is added to the preprogrammed value pH□ of the first line oil pressure calculated in step 311 or S13 according to the following equation. , the first line oil pressure PR to be regulated by the first pressure regulating valve 48 is determined.

...αQ Here, the feedback ring of equation 061 is a so-called P consisting of a proportional action (P action) term and an integral action (l action) term.
KP and TI are proportional gain and integral time, respectively. Further, the correction value determined in step S8 is used for the operation signal Z, but this aQ equation is originally a speed ratio deviation 1e'' in a steady state.
-el/e to match the target deviation value ε.
This is a control formula for determining the line oil pressure PR, and as is clear from FIG. 6, in steady state, the operation signal Z determined according to the formula (2) is applied as is.

The target deviation value C in equation (2) can be obtained through experiments or simulations as a value that provides the best fuel efficiency in the overall driving state of the vehicle, such as driving in the 1O mode, or can be obtained by calculating the fuel efficiency in various driving states. The target deviation value ε is set by various means, such as setting the minimum value of the steady-state deviation required to minimize it as the target deviation value ε. And the first
The line oil pressure Pl is the speed ratio deviation 1 e M″-el/
e is determined so as to match this target deviation value ε.

For example, if the actual first line oil pressure is smaller than it should be, the slope of the oil pressure characteristics shown in FIG.
l/e increases, and 1e*-el/e-ε takes on a positive value. Therefore, the first line oil pressure P j! l (P
An amount proportional to the positive operation signal is added to l, and steps S15 and S16, which will be described later, are executed based on the first line oil pressure PIl obtained in this way, so that the actual first line oil pressure is is increased, the speed ratio deviation 1e*-e l/e is decreased, and is finally made to match the target deviation value ε. Also, if the actual first line oil pressure is larger than the value it should be, the steady-state deviation 1e"-
el/e becomes smaller and l e''-e l/e-ε becomes a negative value. Therefore, the first line oil pressure P l l
An amount proportional to the negative operating signal is subtracted from (P), and the thus determined first line oil pressure P1. By executing steps S15 and S16, which will be described later, based on and the speed ratio deviation l
When e''-e1/e and the target deviation value ε are matched, the actual first line oil pressure becomes the lowest oil pressure value necessary to realize the speed ratio e including the target deviation value ε.

That is, this step 314 is the step Sll
Or the first line oil pressure P R calculated in S13
By correcting r (P) based on the speed ratio deviation le'-el/e, the actual first line oil pressure is controlled to the minimum necessary value. As a result, the drive loss of the pump 42 during steady state is reduced, and the fuel efficiency of the vehicle is improved.

On the other hand, in a quasi-steady state or during shifting when the speed ratio deviation 1e*-el/e becomes large, the correction value ZC, which is smaller than the value of the actual operating signal Z determined by equation (2) above, is used for control as the operating signal. Applies to formula 〇.

For this reason, the speed ratio deviation 1 e II during quasi-steady state or shifting
-e The difference between l/e and the target deviation value ε, that is, the actual operating signal Z, is extremely thick (even if this happens, the first line hydraulic pressure determined by control formula 〇〇 will not become excessive, and the operation The relationship in equation (3) above for modifying signal 2 is:
In this way, even if the speed ratio deviation 1e8-el/e becomes large, it is set in advance using a data map, an arithmetic expression, etc. so that an appropriate first line oil pressure PR+ can be obtained. In this embodiment, step S8 for modifying the operation signal Z corresponds to operation signal modification means.

Then, when the first line oil pressure P1° and the second line oil pressure Pl are determined in this way, next step S1
5 is executed, and the first line hydraulic pressure control value v1 and the second line hydraulic pressure control value v2 for operating the first pressure regulating valve 48° and the second pressure regulating valve 58 are respectively set so that these hydraulic pressures PRt and PI3 are obtained. It is determined. Thereafter, the final step S16 is executed, whereby the first line hydraulic pressure control value Vl+second line hydraulic pressure control value v2° and the speed ratio control value V determined in the step S6 are
. Based on this, drive signals VDI, VD2 and speed ratio signals RAI, RA2 are output.

As a result, the speed ratio e, the first line oil pressure PR, and the second line oil pressure P12 are changed from the speed change control valve 44 to the first pressure regulating valve 4.
8 and the second pressure regulating valve 58, and thereafter, steps 81 and subsequent steps are repeatedly executed.

Thus, in this embodiment, the speed ratio deviation Ie"-el/e between the speed ratio e and the target speed ratio e1 is the target deviation value ε
Since the 1y4th pressure valve 48 is feedback-controlled to match the 1y4th pressure valve 48, the 1st line oil pressure regulated by the 1st gIl pressure valve 48 is the minimum necessary to realize the speed ratio including the target deviation value ε. It is assumed to be a hydraulic pressure value, and the pump 42
The driving loss of the engine 10 and the power loss of the engine 10 are reduced, and the fuel efficiency of the vehicle is improved.

In addition, since the first line oil pressure is regulated based on the speed ratio deviation 1 e'' -e l/e, the first line oil pressure is adjusted in the pressure regulating system including the first pressure regulating valve 48 and the second pressure regulating valve 58 Hydraulic pressure, 2nd line oil pressure - Even if there is a pressure adjustment error or variations in the characteristics of the speed change control valve 44, the actual 1st line oil pressure is always controlled to the minimum necessary oil pressure value. , 1st line hydraulic control value ■
1 and the actual first line oil pressure, the absolute value of the oil pressure does not need to be very accurate. Therefore, it is not necessary to use highly accurate pressure regulating systems and speed change control valves 44 (and the manufacturing costs thereof can be reduced).

On the other hand, in this embodiment, the correction value Zc is applied as the operation signal of the control formula 〇[9 for performing the feedback control, and a sufficient feedback effect can be obtained in accordance with the actual operation signal Z during steady state. , the feedback effect becomes smaller during quasi-steady conditions or when changing gears. Therefore, even during gear shifting when the difference between the speed ratio deviation 1e"-el/e and the target deviation value ε increases, the first line oil pressure P1. does not become excessive, and the first line oil pressure Pit is not excessively increased. This prevents an increase in drive loss of the pump 42 and a decrease in operational maneuverability caused by When the shift value exceeds a certain value Z2, the correction value Z is set to 0, and the feedback effect is completely eliminated. However, in this case, it is not necessary to control the first line oil pressure Pi as strictly as during steady state, so the above-mentioned There is no similar problem even if the darkness of the pressure regulating system is reduced as described above.

Furthermore, in this embodiment, as shown in FIG. 6, since the correction value Zc can be changed continuously, the correction value Zc can be changed in stages.
The first line oil pressure Pi can be controlled smoothly compared to the conventional case where the control format is changed between steady state and gear shift.

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

For example, although the control equation (〇〇) in the above embodiment includes a proportional action term and an integral action term, it is also possible to use a control equation that includes only one of them, or a differential action term.

Furthermore, in the above embodiment, the preprogrammed value P11 (Pl) of the first line oil pressure is obtained in advance, and the final first line oil pressure Pβ1 is determined by adding the feedback term to it. It is also possible to determine the first line oil pressure PJ only by the feedback ring without determining the program value P 12 + (P). In that case, the correction value zc at the time of shifting is set to 0. It is desirable to set it to a positive value.

Further, in the feedback control of the embodiment, the target deviation value ε in the proportional action term and the integral action term is the same value, but it is also possible to set different target deviation values. In particular, the target deviation value of the proportional action term may be 0,
In that case, if feedback control of only proportional action is performed, the control equation will be the preprogrammed value P If t (P)
A hydraulic pressure value proportional to the speed ratio deviation 1 ell-e 1/e is added to , which is also an aspect of feedback control in the present invention.

In addition, in the above embodiment, the speed ratio deviation is 1e1-el/
Although e is used, feedback control can also be performed based on l e'' - e l.

Further, the rotational speed N of the secondary rotating wheel 18. If U is determined, the speed ratio e and the rotational speed N1 of the primary rotating shaft 16 (or the engine rotational speed N,) will have a constant relationship, so the rotational speed N in and the target rotational speed N1.1' By performing feedback control based on the deviation IN5, -N5-1/Ni, ``, the speed ratio deviation l e'' -e
It is also possible to make l/e coincide with the target deviation value ε.

Further, the speed change control valve 44 of the embodiment described above is adapted to simultaneously control the inflow and outflow of the hydraulic oil of both the hydraulic cylinders 26 and 28. The present invention can be similarly applied to other types of hydraulic control devices using control servo valves or the like as speed change control valves.

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

[Brief explanation of the drawing]

FIG. 1 is a configuration diagram of a hydraulic control system for a vehicle belt-type continuously variable transmission, which is an embodiment of the present invention. FIG. 2 is a diagram showing the output oil pressure characteristics of the speed change control valve of FIG. 1. FIG. 3 is a flowchart for explaining the operation of the device on the first page. FIG. 4 is a diagram showing the relationship between the target rotational speed of the primary rotating shaft and the throttle valve opening in the apparatus of FIG. 1. FIG. 5 is a diagram showing a minimum fuel consumption rate curve of the engine of FIG. 1. FIG. 6 shows step S8 of the flowchart in FIG.
FIG. 3 is a diagram showing the relationship between a correction value Zc and an operation signal Z in FIG. 14: Belt type continuously variable transmission 16: Primary side rotating shaft 18 2 Secondary side rotating wheel 20 Knee variable pulley 22: Secondary variable pulley 24: Transmission belt 26: Primary side hydraulic cylinder 28 2 Secondary side hydraulic pressure Cylinder 44: Shift control valve 48: First pressure regulating valve (pressure regulating valve) 94: Controller Pll: First line oil pressure (line oil pressure) Z: Operation signal ゛Z, constant value Z: Correction value Step S8: Operation signal correction Means Applicant: Toyota Motor Corporation Figure 6 Engine Frequency

Claims (3)

[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; In the belt-type continuously variable transmission for a vehicle, comprising a pair of primary hydraulic cylinders and a secondary hydraulic cylinder that respectively change the effective diameters of the pair of variable pulleys, the primary hydraulic cylinder and the secondary hydraulic cylinder are A shift control valve that controls the inflow and outflow of hydraulic fluid in at least one hydraulic cylinder to adjust the speed ratio of the continuously variable transmission, and a pressure regulating valve that regulates the line hydraulic pressure supplied to the shift control valve. The speed change control valve is controlled so that the actual speed ratio of the continuously variable transmission matches the target speed ratio determined according to the driving state of the vehicle, and the actual speed ratio and the target speed ratio are controlled. Feedback control of the pressure regulating valve is performed using the difference between the speed ratio deviation and the target deviation value as an operating signal so that a line oil pressure is obtained to match the speed ratio deviation with a predetermined constant target deviation value. The hydraulic control device is characterized in that an operation signal correction means is provided for correcting the operation signal to a value smaller than the actual value when the operation signal becomes larger than a predetermined constant value. Hydraulic control device for vehicle belt type continuously variable transmission. (2) The hydraulic control device for a belt-type continuously variable transmission for a vehicle according to claim 1, wherein the correction value of the operation signal by the operation signal modification means is continuously changed. (3) The control equation for the feedback control is one in which a feedback term is added to a pre-programmed term for line oil pressure calculated in advance to determine the line oil pressure to be regulated by the pressure regulating valve, and the operation signal correction means is The hydraulic control device for a belt-type continuously variable transmission for a vehicle according to claim 1 or 2, wherein the operating signal is finally set to 0.