JPH023753A - Speed change control device for v-belt type continuously variable speed change gear for vehicle - Google Patents

Abstract

PURPOSE:To provide most excellent fuel consumption by a method wherein the regulated oil pressure of an oil pressure regulating device is controlled to a necessary minimum pressure by means of an engine output and a torque ratio, serving as parameter, and a pressure is increased in a maximum torque ratio and in the vicinity thereof. CONSTITUTION:A line pressure being an output from a regulator valve 61 is controlled to a necessary minimum pressure by means of an oil pressure regulating device 60, a detente valve 64 and a throttle valve 65, interlocked with a throttle opening, and a torque ratio valve 66 interlocked with a torque ratio between input and output pulleys. Meanwhile, a line pressure is increased in a maximum torque ratio and in the vicinity thereof, and is fed to a hydraulic servo on the output side through an oil passage 1 and to the hydraulic servo of a pulley on the input side through a torque ratio control device 80. This constitution enables improvement of durability of a V-belt and fuel consumption.

Description

DETAILED DESCRIPTION OF THE INVENTION [Field of Industrial Application] The present invention relates to a speed change control device for a V-belt type continuously variable transmission for a vehicle.

[Conventional technology]

■The belt-type continuously variable transmission uses an endless metal V-belt invented by Pan d'Orne wrapped around two pulleys. It is characterized in that the V-shaped blocks are arranged without interruption, and power is transmitted between the side surfaces of the pulley and the V-shaped blocks and by the pressing force between the V-shaped blocks.

Conventionally, various proposals have been made for using this endless metal V-belt as a continuously variable transmission. The challenge for a continuously variable transmission is to find a way to accurately change the torque ratio according to the driving conditions.
There are hydraulic control devices to control power transmission according to the torque ratio, and forward/reverse switching mechanisms to prevent shock when shifting from N range to D or R range, but no effective proposals have been made yet. .

FIG. 33 shows a conventional continuously variable transmission for vehicles proposed in Japanese Patent Application Laid-Open No. 54-159730. input shaft a
is provided with an input-side pulley consisting of a fixed flange b and a movable flange C, and an output-side pulley consisting of a fixed flange e and a movable flange f is provided on the output shaft d. A belt g is stretched around the shaft, and power is transmitted from the input shaft a to the output shaft d. This control device for a continuously variable transmission for a vehicle is provided with a regulator valve k that regulates the hydraulic pressure generated by a pump j to a predetermined line pressure, and the line pressure regulated by the regulator valve k is passed through an oil path i. is supplied to the movable flange f of the output pulley to give the output pulley a force to clamp the bell (g), and the oil path connected to the movable flange C of the input pulley is selectively supplied by the torque ratio control valve 2. Hydraulic pressure is supplied to and discharged from the movable flanges c and f, thereby moving the movable flanges c and f.

Moreover, at one end of the spool m of the torque ratio control valve l,
Fluid pressure proportional to the rotational speed of the input shaft a acts on the pitot tube n, while on the other end of the spool m, pressure due to the rotation of the cam p linked to the movement of the throttle pedal acts on the lever q and the spring r. It acts through. Therefore, the movable flange C of the input pulley due to the torque ratio control valve i
Hydraulic pressure is supplied and discharged according to the rotational speed of the input shaft a and the movement of the throttle pedal. Furthermore, fluid pressure proportional to the rotational speed of the input shaft a acts on one end of the spool S of the regulator valve through the pitot tube n, while the input pulley is affected by the axial movement of its movable flange C. A detection rod t is provided which is displaced in conjunction with the spool S, and the pressure of the detection rod t acts on the other end of the spool S via a lever U and a spring V. Therefore, the pressure regulating action on the regulator valve is configured to be performed in accordance with the rotational speed of the input shaft a and the torque ratio of the continuously variable transmission.

[Problem to be solved by the invention]

By the way, in controlling the torque ratio mentioned above, the difference between this type of continuously variable transmission and a general automatic transmission is that the continuously variable transmission transmits power using a small frictional resistance between the metal pulley and the metal block. To do this, high hydraulic pressure is required, and the engine is idling (throttle opening fully closed).
Also, a relatively high hydraulic pressure is required to prevent the V-belt from slipping during engine braking.

However, in the control of the torque ratio of the conventional continuously variable transmission mentioned above, the setting is such that no slip occurs with respect to the maximum engine output torque for each throttle opening, so in normal driving, the maximum engine output torque is Since the output torque is smaller than the output torque, a force greater than the required belt clamping force is applied, which impairs the belt's durability and increases the oil pump drive torque, reducing fuel efficiency. There is a problem.

Therefore, it is conceivable to control the clamping force so that it does not cause slip in relation to the output torque at the target input rotation speed during normal driving, but for example, when starting, the output torque at the target input rotation speed In a region where the output torque is high, the line pressure is insufficient, that is, the belt clamping force is insufficient, and the belt slips.

The present invention solves the above problems, and its main purpose is to control the line pressure to the minimum necessary pressure using the engine output torque and the torque ratio between the input and output pulleys as parameters. The objective is to improve durability and fuel efficiency.

Another object of the present invention is to generate appropriate clamping force on the V-belt and prevent belt slippage by increasing the line pressure in conditions of maximum torque ratio, such as when starting or climbing a mountain. It is to be.

[Means to solve the problem]

For this purpose, the present invention provides fixed flanges (311, 321)
, an input end pulley (3I) and an output pulley (32) having a movable flange (312, 322) and a hydraulic servo (313, 323) that displaces the movable flange;
V-bell 1-(33) stretched between these two pulleys,
A hydraulic pressure regulating device (60) that regulates the hydraulic pressure generated in the hydraulic pressure source (52), and an input side pulley by selectively supplying the hydraulic pressure regulated in the hydraulic pressure regulating device to the hydraulic servo (313, 323). (31) and the torque ratio fIIIg that makes the effective diameter of the output pulley (32) variable
I device (80), and the hydraulic pressure adjustment device! (60)
The hydraulic pressure regulated in the engine is controlled to the minimum necessary pressure by using the output torque of the engine and the torque ratio between the input and output pulleys as parameters, and the pressure is increased at and around the maximum torque ratio. do.

[Action and effect of the invention]

In the present invention, in order to run the vehicle with low fuel consumption, it is necessary to bring the line pressure supplied to the hydraulic control circuit close to the necessary minimum, and in the continuously variable transmission, the line pressure is applied to the input pulley 31 and the output side. Each hydraulic servo of the pulley 32 is defined by a hydraulic pressure that allows torque to be transmitted without causing the V-belt 33 to slip. When the vehicle starts, it is impossible to operate the engine at the best fuel efficiency within the range of torque ratios that can be achieved by both pulleys. The line pressure shown by the broken line is approximately 20% larger.

To achieve this, the line pressure, which is the output of the regulator valve 61, is adjusted by the oil pressure adjustment device 60 as follows by changing the throttle opening θ and the torque ratio between the input and output shafts.

For example, in the D range, the manual valve 62
Only the oil passage 1 communicates with the oil passage 1, and the oil passage 4 and the oil passage 5 are depressurized. At this time, in the shift control mechanism 70,
When the shift control solenoid valve 74 is in the OFF state, the oil chamber 71
If line pressure is supplied to spool 71
2 is located on the right side, the oil passage 3 and the oil passage 13 are connected, and the line pressure supplied to the oil passage 3 acts on the hydraulic servo 49 of the forward multi-disc clutch 45 through the oil passage 13. , the vehicle is ready to move forward.

When the torque ratio T is large (1×T≦j4), the pressure in the oil passage 9 is exhausted from the drain boat 666 as shown in FIG. 5 (-C), and therefore the throttle pressure is as shown in FIG. 8 (D). be done. However, since the port 663 is opened and the oil passage 1 and the oil passage 6 communicate with each other, the throttle opening θ is within the range of 0≦θ≦01%, and the spool 641 of the detent valve 64 is moved as shown in FIG. As shown in the figure, while the oil passage 6 is closed by the spool 641 and pressure is discharged from the manual valve 62 through the oil passage 5, the throttle opening θ is 01%. When the value is 05100%, Figure 4 (
As shown in B), the spool 641 moves to the right in the figure, and the oil passage 6 and the oil passage 7 communicate with each other, and detent pressure is generated in the oil passage 7. As a result, the line pressure is reduced to the area (w) in Figure 9 and the first line pressure.
As shown in the θ diagram (January), the characteristic changes stepwise at θ-01%.

Therefore, according to the present invention, the durability and fuel efficiency of the V-belt can be improved by controlling the line pressure to the minimum necessary pressure using the output torque of the engine and the torque ratio between the input and output pulleys as parameters.

In addition, by increasing the line pressure at and near large torque ratios such as when starting or climbing a hill, V
It is possible to increase the input rotation speed by generating a clamping force on the belt, and it is possible to eliminate the drawbacks of the method of controlling the line pressure to the minimum necessary pressure.

Note that the numbers added to the above-described configurations are for comparison with the drawings to facilitate understanding, and the configurations are not limited thereby.

〔Example〕

Embodiments of the present invention will be described below with reference to the drawings.

FIG. 1 is a schematic diagram of a V-belt type continuously variable transmission for a vehicle.

100 is an engine, 102 is a carburetor, 20 is a transmission device provided between the engine 100 and the drive side axle, and a fluid coupling 21 connected to the output side 101 of the engine, and a fluid coupling 21 connected to the fluid coupling 21. V-belt continuously variable transmission 30. A planetary gear transmission 4 for forward/reverse switching connected to the output shaft 26 of the continuously variable transmission 30
0, it is constituted by a continuously variable transmission which basically consists of a gear mechanism 23 connected to an output shaft 47 of the planetary gear transmission 40.

The fluid coupling 21 is connected to the output shaft 101 of the engine.
a pump impeller 211 connected to a turbine runner 21 connected to a fluid coupling output shaft 214;
This is a well-known method consisting of 2. Note that other fluid torque converters or mechanical clutches may be used instead of the fluid coupling.

The V-belt continuously variable transmission 30 includes an input pulley 31 connected to an output shaft 214 of a fluid coupling, and an output shaft 26 of the V-belt continuously variable transmission arranged parallel to the input pulley 31. The output pulley 32 is connected to the output pulley 32, and the belt 33 is stretched between the two pulleys.

The input pulley 31 has a fixed flange 311 connected to the output shaft 214, and a movable flange 312 provided opposite to the fixed flange 311 to form a V-shaped space. It is provided so as to be movable in the axial direction by a hydraulic servo 313.

The output pulley 32 includes a fixed flange 321 connected to the output shaft 26 of the continuously variable transmission 1130, and the fixed flange 32.
1 and a movable flange 322 provided so as to form a V-shaped space, and the movable flange 322 is provided so as to be movable in the axial direction by a hydraulic servo 323.

The planetary gear transmission 40 for forward/reverse switching includes a sun gear 41,
Ring gear 43, these sun gear 41, ring gear 43
The sun gear 41 is connected to the output shaft 26 of the continuously variable transmission, and the carrier 46 is a planetary gear for forward/reverse switching. It is connected to an output shaft 47 of the transmission 40. The sun gear 41 and the carrier 46 are removably connected by a multi-disc clutch 45, and the ring gear 43 is removably connected to a transmission case 400 by a multi-disc brake 42.

In this planetary gear transmission 40 for forward/reverse switching, when hydraulic pressure is supplied to the hydraulic servo 49, the multi-plate clutch 45 engages and the rotation of the output shaft 26 of the continuously variable transmission continues to the planetary gear transmission for forward/reverse switching. The signal is transmitted to the output shaft 47 of the machine 40 to enable forward running. Further, when hydraulic pressure is supplied to the hydraulic servo 48, the multi-disc brake 42 is engaged and the ring gear is fixed, so that the output shaft 47 is connected to the output shaft 2 of the continuously variable transmission.
It rotates in the opposite direction to the rotation of No. 6 to enable a backward running state.

The reduction gear mechanism 23 is for compensating for the fact that the speed change range obtained by the V-belt type continuously variable transmission 30 is lower than the speed change range achieved by a normal vehicle transmission. It decelerates and increases torque. The output shaft of the reduction gear mechanism 23 is connected to the differential gear 22, and performs final reduction at a reduction ratio of 3.727, for example.

FIG. 2 shows a hydraulic control circuit for the V-belt continuously variable transmission shown in FIG.

The hydraulic control circuit includes a hydraulic source 50, a hydraulic adjustment device 60, and an N-
Shift control mechanism 7 that alleviates the impact during D and N-R shifts
0, and a torque ratio control device 80.

The hydraulic power source 50 supplies hydraulic oil pumped up from an oil reservoir through an oil strainer 51 by a pump 52 driven by an engine to a regulator valve 61 through an oil passage 11 to which a relief valve 53 is attached.

The hydraulic adjustment device 60 includes a manual valve 62 that is manually operated by a shift lever (not shown), a detent valve 64 and a throttle valve 65 that output detent pressure and throttle pressure according to the throttle opening θ of the carburetor 102, and an output. A torque ratio valve 66 that operates in conjunction with the movable flange 322 of the side pulley 32 and supplies line pressure to the detent valve 64 according to the amount of displacement of the movable flange 322, and discharges pressure from the output hydraulic pressure feedback oil passage 9 provided in the throttle valve 65, and a hydraulic pressure 1150. It is comprised of a regulator valve 61 that regulates the hydraulic pressure supplied from and supplies it to the hydraulic servo 323 as line pressure.

The manual valve 62 is set so that the spool 621 is set to P, 8%N1D, and L positions as shown in FIG. It is table! As shown in the figure, the oil passage 1 to which line pressure is supplied is connected to the output oil passages 3 to 5.

table ! In Table 1, 0 indicates a state of communication with oil passage 1, and x indicates that oil passages 3 to 5 are in a discharged pressure state.

The regulator valve 61 includes a spool 611 and a regulator pulp plunger 612 that controls the spool 611 by inputting detent pressure and throttle pressure. Adjust the line pressure and output the line pressure from the output port 16 to the oil path l. Oil discharged from boat 614 is supplied to fluid couplings, oil coolers, and lubrication points via oil line 12.

The detent valve 64 includes a spool 641 that moves in conjunction with the throttle opening θ of the butterfly valve of the capretor 102 as shown in FIG.
01, the oil passage 5 and the detent pressure output oil passage 7 connected to the input port 16' provided in the regulator valve 61 are communicated as shown in FIG.
When it is 100%, the oil passage 7 and the oil passage 6 which connects the detent valve 64 to the torque ratio valve 66 are communicated as shown in FIG. 4(B).

The throttle valve 65 is arranged in series with the spool 641 of the detent valve via a spring 645, and includes a spool 651 with a spring 652 installed in front of the other side, and throttle opening transmitted via the spool 641 and the spring 645. Due to the action of the spool 6510 that moves according to the fluctuation of the degree θ, the boat 65 that communicates with the oil path 1
3 and outputs throttle pressure to the throttle pressure output oil passage 8 which communicates with the input port 18 provided in the regulator valve 61. The spool 651 branches from the oil path 8 and outputs to a land 656 and a land 657 having a larger pressure receiving area than the land 656 via output oil pressure feedback oil paths 9 and 10 provided with orifices 654 and 655, respectively. Receiving hydraulic feedback.

The torque ratio valve 66 has a spool 6 linked to the movable flange 322 of the output pulley 32 via a connecting rod.
62, and the amount of movement of the movable flange 322 is l, ≦L.
≦I! 4 (Torque ratio T is t.

≧T≧j+), the spool 662 is located on the left side of the figure as shown in FIG. At the same time, the output oil passage 6 to the detent valve 64 is communicated with a drain boat 665 to discharge pressure. The amount of movement of the movable flange 322 is the first set value! 3
When smaller (,18≦L<z, (ts≧Tit,), as shown in FIG. 666, and the oil passage 9 is depressurized.

When the amount of movement of the movable flange 322 is smaller than the second set value 1 and 0≦L≦lx (t4 ≧Tit,), the spool 662 is located on the right side of the figure as shown in FIG. 5(C). , a boat 663 connected to the oil passage 1 and the oil passage 6 communicate with each other, and line pressure is supplied to the oil passage 6.

The shift control mechanism 70 includes a shift control valve 71 including a spool 712 having a spring 711 installed in front of one end and receiving line pressure from an oil chamber 713 provided at the other end;
Orifice 7 provided in the oil passage l that supplies heline pressure
2, a pressure limiting valve 73 installed between the orifice 72 and the oil chamber 713, and a solenoid valve 74 that adjusts the oil pressure of the oil chamber 713 by being controlled by an electric charge m circuit (described later).

When the solenoid valve 74 is turned on to open the drain boat 741 and evacuate the oil chamber 713, the shift control valve 71
The spool 712 is moved to the left in the figure by the action of a spring 711, and is connected to the oil passage 13 that connects to the hydraulic servo 49 that operates the multi-disc clutch 45 of the planetary gear transmission 40 and the hydraulic servo 48 that operates the multi-disc brake 42. The communicating oil passages 14 are connected to the drain boats 714 and 715, respectively, to discharge pressure, and the multi-disc clutch 45 or the multi-disc brake 42 is released.

When the solenoid valve 74 is off, the drain boat 741 is closed, and the spool 712 is located on the right side in the figure due to the line pressure supplied to the oil chamber 713, connecting the oil passages 3 and 4 to the oil passage 13, respectively. and the oil passage 14, and engages the multi-disc brake 42 or the multi-disc clutch 45.

In this embodiment, the shift control valve 71 is provided with an oil chamber 717 and an oil chamber 716 that feed back the output oil pressure of the oil passage 13 and the oil passage 14, and this reduces the rise of the output oil pressure and controls the multi-disc clutch 45 and the multi-disc brake 42. This prevents shock when mating.

The torque ratio control device 80 includes a torque ratio control valve 81, orifices 82 and 83, and a downshift solenoid valve 84.
, and an upshift solenoid valve 85.

The torque ratio control valve 81 includes a spool 812 with a spring 811 installed in front of it on one side, oil chambers 815 and 816 at both ends to which line pressure is supplied from the oil passage 1 through orifices 82 and 83, respectively, and line pressure is supplied. Oil road 1
A human-powered boat 817 whose opening area increases and decreases according to the movement of the spool 812, and an output boat 818 which communicates with the hydraulic servo 313 of the input pulley 31 of the V-belt continuously variable transmission 30 via the oil passage 2. , a drain boat 814 that evacuates the oil chamber 819 as the spool 812 moves, and a drain boat 813 that evacuates the oil chamber 815 as the spool 812 moves.
Equipped with The downshift solenoid valve 84 and the upshift solenoid valve 85 are respectively attached to the oil chamber 815 and the oil chamber 816 of the torque ratio control valve 81,
Both are operated by the output of an electric control circuit to be described later, and evacuate the pressure in the oil chamber 815 and the oil chamber 816, respectively.

FIG. 6 shows the solenoid valve 74 of the shift control mechanism 70 and the torque ratio control device 80 in the hydraulic control circuit shown in FIG.
The configuration of an electric control circuit 90 that controls the downshift solenoid valve 84 and the upshift solenoid valve 85 is shown.

901 is a shift lever switch that detects whether the shift lever is shifted to P, R, N, D, or L; 902 is a rotation speed sensor that detects the rotation speed of the input pulley 31; 903 is a vehicle speed sensor; 904 is a throttle sensor that detects the throttle opening of the carburetor or the amount of depression of the accelerator pedal, 905 is a speed detection processing circuit that converts the output of the rotation speed sensor 902 into voltage, and 906 is a vehicle speed that converts the output of the vehicle speed sensor 903 into voltage. Detection circuit, 907 is a throttle opening detection processing circuit that converts the output of the throttle sensor 904 into voltage, 908 to 911 are human interfaces for each sensor, 912 is a central processing unit (CPU), 913 is a solenoid valve 74.84.85
914 is a read-on memory (ROM) that stores programs to control the program and data necessary for control; 914 is a random access memory (RAM) that temporarily stores input data and parameters necessary for control; 915 is a shock; 916 is a random access memory (RAM) that temporarily stores input data and parameters necessary for control; is an output interface, 917 is a solenoid output driver, and output interface 9
16 is converted into the operating output of the amplifier shift solenoid valve 85, the downshift solenoid valve 84, and the shift control solenoid valve 74. Input interfaces 908 to 911, CPU 912, ROM 913,
The RAM 914 and the output interface 916 are connected through a data bus 918 and an address bus 919.

Next, the operation of the hydraulic pressure adjusting device 60, which is a feature of the present invention and is comprised of a torque ratio valve 66, a detent valve 64, a throttle valve 65, a manual valve 62, and a regulator valve 61, will be explained.

The hydraulic oil supplied to the hydraulic control circuit is supplied from a pump 52 driven by the engine, and if the line pressure is high, power consumption by the pump 52 increases accordingly. Therefore, in order to run a vehicle with low fuel consumption, hydraulic pressure is required! 1I
It is necessary to bring the line pressure supplied to the J11 circuit close to the necessary minimum, and in the continuously variable transmission, the line pressure can be adjusted so that each hydraulic servo of the input pulley 31 and the output pulley 32 does not cause the V belt 33 to slip. It is defined by the hydraulic pressure that can transmit torque.

When the engine is operated in a state that provides the best fuel efficiency, the minimum necessary line pressure for changes in the torque ratio T between the input and output shafts is shown by the solid line in FIG. 7 using the throttle opening θ as a parameter. When starting the vehicle, it is impossible to operate the engine at the best fuel efficiency within the range of torque ratio that can be achieved by both pulleys, so as shown by the dotted line, it is approximately 20% lower than the best fuel efficiency characteristic curve shown by the solid line above. It is desirable to set the line pressure to the line pressure shown by the large broken line, and it is also desirable to set the line pressure to a higher line pressure characteristic shown by the dashed line even at throttle opening θ#0 during engine braking.

In this embodiment, the line pressure, which is the output of the regulator valve 61, is controlled by the hydraulic pressure adjusting device 60 to the manual valve 6.
It is adjusted as follows by changing the shift position (L, D, N, RlP) of 2, the throttle opening θ, and the torque ratio between the input and output shafts.

(Position D) As shown in Table 1 above, in the manual valve 62, only the oil passage 3 communicates with the oil passage 1, and the oil passage 4 and the oil passage 5 are depressurized. At this time, in the shift control mechanism 70, the shift control solenoid valve 74 is in the OFF state, and the oil chamber 7
If line pressure is supplied to spool 7
12 is located on the right side, the oil passage 3 and the oil passage 13 are connected, and the line pressure supplied to the oil passage 3 acts on the hydraulic servo 49 of the forward multi-disc clutch 45 through the oil passage 13. , the vehicle is ready to move forward.

■ When the torque ratio T is t, ≦T≦t2.

As shown in FIG. 5(A), the torque ratio valve 66
The port 663 connected to the drain port 665 is closed, and the oil passage 6 is communicated with the drain port 665 to exhaust pressure.

As a result, regardless of the throttle opening θ, the oil passage 7
No detent pressure is generated, and the throttle valve 65 is
The port 664 of the torque ratio valve 66 that communicates with the oil passage 9 is closed, and the spool 651 receives feedback pressure from the land 657 as well as the land 656, so that the throttle opening θ changes as shown in FIG. A throttle pressure having the characteristics shown in is outputted to the regierator valve plunger 612 of the regulating valve 61 via the oil passage 8. As a result, the line pressure output from the regulator 61 is in the region (f) of FIG.
The result is as shown in Figure 0 (E).

■ When the torque ratio T is t, <'r≦t.

5th rj! As shown in J(B), the torque ratio valve 66 closes the boat 663, and the oil passage 9 and the drain boat 66
6. The oil passage 6 is also depressurized through the boat 665. Therefore, no detent pressure is generated, and the throttle pressure is reduced to the land 65 of the spool 651 when the oil passage 9 is exhausted.
6 increases by the amount that the feedback pressure is no longer applied, and is represented by the characteristic curve shown in FIG. 8(d). The line pressure at this time is area (L) in Figure 9 and area (G) in Figure 10.
It has the characteristics shown below.

■ When the torque ratio T is t, <7'≦t4.

As shown in FIG. 5(C), the oil passage 9 has a drain 1666
Therefore, the throttle pressure is the same as in ① above.
This is shown in figure (d). However, since the port 663 opens and the oil passage 1 and the oil passage 6 communicate with each other, the throttle opening θ becomes O.
≦θ≦θ, within the range of 2, and while the spool 641 of the detent valve 64 is on the left side of the figure as shown in FIG. 4(A), the oil passage 6 is closed by the spool 641 and the oil passage is 7 is exhausted from the manual valve 62 via the oil passage 5, but when the throttle opening θ is 03% to 05100%, the spool 641 is moved to the right side in the figure as shown in FIG. 4(B). The oil passage 6 and the oil passage 7 communicate with each other, and detent pressure is generated in the oil passage 7. As a result, the line pressure is reduced to (
As shown in area (w) and (s) in Figure 10, θ=θ1%
The characteristic changes stepwise.

(L position) At the manual valve 62, the oil passage 5 communicates with the oil passage l.

Oil passage 3 and oil passage 4 are equivalent to the D position.

■ When the torque ratio T is tl≦T≦1t.

When the throttle opening degree θ is 0≦θ≦01%, the oil passage 5 and the oil passage 7 communicate with each other at the detent valve 64, and detent pressure is generated to push up the throttle plunger and generate high line pressure. Throttle opening θ is 81%
5100%, the oil passage 7 is the oil passage 6 and Fig. 4 (B)
As shown in the figure, the pressure is exhausted through the drain port 665 of the torque ratio valve, so no detent pressure is generated, and the throttle pressure is the same as in the D position. Therefore, the line pressure is the first
The characteristics are shown in (R) in Figure 1.

■ When the torque ratio T is t, <T≦t.

The difference from the above (■) is that in the torque ratio valve 66, the oil passage 9
is communicated with the drain port 666 and is exhausted, and the throttle pressure output from the throttle valve 65 to the regulating valve 61 via the oil passage 8 increases, and as a result, the line pressure reaches (h) in Fig. 11. It is expressed by the characteristic curve as shown.

■ When the torque ratio T is t, <“l’≦t4.

The torque ratio valve 66 communicates the oil passage 6 with the oil passage 1, and the oil passage 9 is depressurized from the drain port 666. Since line pressure is supplied to both the oil passage 6 and the oil passage 5, the detent valve 64 outputs the detent pressure regardless of the throttle opening, and the regulator valve inputs the detent pressure and the same throttle pressure as in (2) above. 61 outputs the line pressure shown in FIG. 11 (N).

(R position) In the manual valve 62, oil passage 4 and oil passage 5 are oil passage 1.
The oil passage 3 is depressurized. At this time, in the shift control mechanism 70, the shift control solenoid 74 is turned off.
When line pressure is being supplied to the oil chamber 713 in the F state, the spool 712 is positioned to the right, so that the oil passage 4 and the oil passage 14 are communicated with each other, and the line pressure supplied to the oil passage 4 is is supplied to the hydraulic servo 48 of the reverse multi-disc brake 42 through the oil path 14, and the vehicle enters the reverse state.

(P position and N position) Since the oil passages 3.4 and 5 are both exhausted in the manual valve 62, the line pressure that is the output of the regulator valve 61 is the same as in the D position.

In this line pressure adjustment, the manual valve 62 is
When shifting to each shift position of P, the line pressure when the torque ratio T is in the range of t2<T≦t4 is set low with the throttle opening θ1% or less as shown in the characteristic curve (S) in Figure 10. The reason for this is that if the line pressure is set high in operating conditions such as idling where the throttle opening θ is small and the pump discharge amount is low, the line pressure will increase when the oil temperature is high and there is large oil leakage from various parts of the hydraulic circuit. This is to prevent this from happening, since it becomes difficult to maintain the pressure, and furthermore, the oil temperature further increases due to a decrease in the amount of oil supplied to the oil cooler, which tends to cause trouble.

Further, when the manual valve 62 is shifted to the R and R shift positions, the torque ratio T is 1.5 as shown in the characteristic curves (H) and (R) of FIG. ≦T≦1. The reason why the line pressure is set high under the operating conditions where the throttle opening θ is 0.1% or less is that a relatively high oil pressure is required during engine braking even when the throttle opening is low.

In this way, as shown in FIG. 9, by bringing the line pressure closer to the minimum necessary oil pressure shown in FIG. 7, the power loss caused by the pump 52 can be reduced, and fuel efficiency can be improved.

Next, the shift control mechanism 70 and torque ratio control device 80 controlled by the electric control circuit 90 explained in FIG.
The operation will be explained with reference to the program flowchart shown in FIG.

After the throttle opening θ is read by the throttle sensor 904 (step 921), the shift lever position is determined by the shift lever switch 901 (step 922). As a result of the determination, if the shift lever is in the P position or N position, the solenoid valves 84 and 85 are activated by the P position or N position processing subroutine shown in FIG.
Both are set to 0FFL (step 931), and the P or N state is stored in the RAM 914 (step 932).

As a result, the input pulley 31 is brought into a neutral state. When the shift lever changes from the P position or the N position to the R position, and when it changes from the N position to the D position, shift shock control is applied to alleviate the shift shock associated with the N-R shift and N-D shift, respectively. Processing is performed (steps 940, 950).

This shift shock control processing will be explained in detail below.

First, the shift control mechanism 70 includes the electric control circuit 90 described above.
By the action of the solenoid valve 74 controlled by the output of It has the function of keeping the upper limit of the hydraulic pressure supplied to the hydraulic servos 48 and 49 below a set value, and limits the holding pressure of the clutch and brake.

In this embodiment, as shown in FIG. 14, the pressure receiving area of the land provided on the spool 712 of the shift control valve 71 is
In order from the left side of the figure, st, s, , s, SN, the elastic force of the spring 711 is F, and the oil pressure of the oil chamber 713 is P. The hydraulic pressure PC and the hydraulic pressure P supplied to the hydraulic servo 48 of the multi-disc brake 42 that is engaged during reverse movement are given as follows from equations 0 and 0, which are hydraulic balance equations for the shift control valve 71, respectively.

When moving forward P s ・S t -P c-3t +F st
-■Pc-(St/S□)Ps (Fs+/S8
) When moving backward Ps -5I=Pb (St S,)"Fs+
■Ph −(s+/(st st))Ps−(F*
+/(St Sri) Further, if the pressure receiving area of the valve body 731 inserted in the pressure limiting valve 73 is SN, and the elastic force of the spring 732 installed on the valve body 731 is F,2, The pressure limiting valve 73 is set to P according to the hydraulic balance type formula
, operates at a maximum pressure P 411sit.

P 11m1t XS 2-F B
■P 11m1t =F sz/Sz At this time, the maximum pressures Pc1i■it%P and Hsit of PC and Pb are limited according to the 0th equation and the 0th equation.

When moving forward P c l1sit = (S +/s x
) P 11m1t-F s+/S t
■When going backwards Ph 11m1t -(S +/, (S + -3*
) ) P 11m1t (P sl/C5t S
z ) ) ■The solenoid valve 74 generates a solenoid pressure P in the oil chamber 713 according to the duty (%) given by the following equation.

Deity (Solenoid ON time in one cycle/Solenoid operation period) ), and the pulse that gradually becomes smaller is called the 14th pulse.
This is done by adding it to the shift control solenoid valve 74 shown in the figure. In this way, the shift control solenoid valve 7
By duty-controlling the shift control valve 71, a hydraulic pressure P3 adjusted according to the duty is generated in the oil chamber 713 of the shift control valve 71.

The solenoid pressure P shown in FIG. 17 is the shift control valve 71.
The hydraulic pressure Pc or P supplied to the hydraulic servo 48 or 49 shown in FIG.

is obtained.

When mitigating locking shock during N-D shift and N-R shift, hydraulic servo 48 or hydraulic servo 4
The rise of the oil pressure P or Pc supplied to the pump 9 is rolled as shown in the oil pressure characteristic curve shown in Fig. 16, and in the figure, AC
The engagement of the multi-disc clutch 45 or the multi-disc brake 42 between the two ends is completed. In this way hydraulic servo 48 or 49
Solenoid valve 7 for controlling the hydraulic pressure supplied to
Shift shock control processing 940.
A program flowchart of 950 is shown in FIG.

Figure 19 shows each parameter K of the waveform diagram shown in Figure 15.
, L", and M". In step 941, it is determined whether or not the PLUG is on during shock control processing, and the PLUG is
If UG is on, shift shock control processing is in progress and the process proceeds to step 946; if PLUG is not on, R is turned on to start shift shock control processing.
By comparing the shift lever position stored in the AM914 with the current shift lever position, it is determined whether the shift lever has changed from the P position or the N position to the R position (step 942) and from the N position to the D position. A determination is made as to whether or not there is a change (step 943). If any change has occurred, in steps 944 and 945 the corresponding parameters K'', L'', M
PLU that indicates that the parameter K is set to O and shock control processing is performed.
G is turned on (step 955). If no change has occurred, the process returns and no shock control processing is performed.

In step 946, it is determined whether the parameter K that determines the end of the cycle K" is larger than O. If K is not thicker than O, K is K"-1 and L is L".
, L'' is set as L''-M'' (step 947), and in step 948 it is determined whether L is less than or equal to θ, and if L is less than or equal to 0, the process proceeds to step 951, and if L is less than or equal to θ , assumes that all shock control processing has been completed and turns off the PLUG.In step 946, the PLUG is
When the parameter Kfr<O is greater than O, K-1 is set to K (step 950), and then it is determined whether the parameter L, which determines the end of the on time in one cycle K, is O (step 950). 951). L is 0
If so, a command to turn off the solenoid valve 74 is issued (step 952), and if L is not O, a command to turn on is issued (step 953), after which L-1 is set to L and the process returns.

Similar shift shock control processing can also be performed using programmable timer 920 shown in FIG.

Next, referring back to FIG. 12, shift control, which is a feature of the present invention, will be explained.

Next to the N-D shift shock control process 950, the actual input pulley rotation speed N is read by the input pulley rotation speed sensor 902 (step 923), and then the throttle opening degree θ read in step 921 is 0.
It is determined whether or not (step 924), and when θ≠0, a subroutine 960 is executed to set the input pulley target rotation speed N1 to the input pulley rotation speed for the best fuel economy, and when θ
When the throttle is fully closed at =O, in order to determine the necessity of engine braking, it is determined whether the shift lever is set to the D position or the L position (
Step 926) When the shift lever is set to the D position, a D position engine brake processing subroutine 970 is executed, and when the shift lever is set to the L position, an L position engine brake processing subroutine 980 is executed. Then, the input pulley target rotation speed N1 is set to a value suitable for each.

A subroutine 960 for setting the input pulley target rotation speed N1 described above to the input pulley rotation speed for the best fuel efficiency will be explained.

Generally, in order to operate the engine at the best fuel efficiency,
It is preferable to operate according to the best fuel efficiency power line shown by the broken line in FIG. In this figure 20, the horizontal axis is the engine speed (
rpm), the vertical axis is the torque of the engine output shaft (kg-m)
The best fuel efficiency power line can be obtained as follows. That is, from the engine's equal fuel consumption rate curve (in g/p s - h) shown by the solid line in Fig. 20 and the equal horsepower curve (in pS) shown by the two-dot chain line, the equation at point A in the figure can be determined. If the fuel consumption rate is Q (g/pS-h) and the horsepower is P (p, s), then at point A, 5=QXP (g/h) of fuel will be consumed per hour. By determining the fuel consumption per hour S at all points on each equal horsepower curve, the point where S is the minimum on each equal horsepower curve can be determined, and by connecting these points, the best value for each horsepower can be determined. The best fuel efficiency power line indicating the engine operating state resulting in fuel efficiency can be obtained.

However, as in this embodiment (the engine 100 and the fluid communication a
T! When the fluid coupling 21 is combined with I, the engine output performance ratio' line at the throttle opening θ shown in Fig. 21 and the fluid coupling performance curve shown in Fig. 22 are obtained using the same method. , the best fuel consumption fluid coupling output line can be found on the fluid coupling output performance curve as shown in FIG. 24 from the engine equal fuel consumption rate curve shown in FIG. 23. Figure 25 is the 24th
The best fuel efficiency fluid coupling output line shown in the figure is replaced with the relationship between throttle opening and fluid coupling output rotation speed. In the continuously variable transmission device of this embodiment, this fluid coupling output rotation speed directly becomes the input pulley rotation speed N.

For this purpose, the input pulley target rotation speed N1 shown in FIG.
In the subroutine 960 for setting the best fuel efficiency input pulley rotation speed N, the best fuel efficiency input pulley rotation speed N corresponding to the throttle opening degree θ shown in FIG. 25 is stored in advance in the ROM 913 as data from the throttle opening θ , Set the data address (step 961), read out the best fuel efficiency input pulley rotation speed N from the set address (step 962), read out the best fuel efficiency input pulley rotation speed corresponding to the read throttle opening θ N, is set to the input pulley target rotation speed N1 (step 963).

Next, the engine brake processing subroutine 9 in FIG.
70,980 will be explained.

The engine brake processing subroutine 970 at position D is as follows:
As shown in FIG. 27, the vehicle speed V is detected by the vehicle speed sensor 903.
is read (971), and at that point the acceleration α is calculated (972), and then it is determined whether the acceleration α is an appropriate acceleration A for the vehicle speed (973), and if the acceleration α is greater than the acceleration A. In this case, the input pulley target rotation speed N0 should be set to a value larger than the current input pulley rotation speed N in order to downshift (974), and when the acceleration α is not larger than the acceleration A, the input pulley target rotation speed N0 should be set to a value larger than the current input pulley rotation speed N.
0 is set to the best fuel consumption input side pulley rotation speed N corresponding to the throttle opening degree θ.975) Return. The relationship between vehicle speed and appropriate acceleration A is determined in advance by experiment or calculation for each vehicle, as shown in FIG.

The engine brake processing subroutine 970 for the L position is as follows:
As shown in FIG. 29, the vehicle speed V is detected by the vehicle speed sensor 903.
is read (981), and then the current torque ratio T is calculated from the vehicle speed V and the input pulley rotation speed N using the following formula (9
82).

T= (N/V)xk k is a constant determined from the reduction ratio of the reduction gear mechanism 23 inside the transmission, the final reduction ratio of the vehicle, the tire radius, etc. 0 Then the current torque ratio T is 983. To determine whether the torque ratio T0 is smaller than that at which safe and proper engine braking can be obtained. ),) When the torque ratio T is smaller than the torque ratio T1, the input pulley target rotation speed N1 is set to a value larger than the current input pulley rotation speed N in order to downshift (984), and the process returns. If the torque ratio T is not smaller than the torque ratio T1, the input pulley target rotation speed N1 is set to the current input pulley rotation speed N (985), and the process returns. The torque ratio T11 that provides safe and appropriate engine braking for the vehicle speed is determined in advance through experiments or calculations for each vehicle, as shown in FIG.

When the input pulley target rotation speed N1 is set as described above, in FIG. ), if NO, issue a command to operate the downshift solenoid valve 84 and proceed to step 928.
), when N>N”, amplifier shift solenoid valve 85
When N=N'', an OFF command is issued for both solenoid valves 84 and 85 (step 920).

The torque ratio control device 80 controls the increase or decrease in the gear ratio between the input and output pulleys by comparing the rotational speed of the input pulley with the best fuel efficiency obtained in FIG. 17 and the actual rotational speed of the input pulley. This is done by operating two solenoid valves 84 and 85 provided in the control device 80, so that the actual input pulley rotation speed is brought to the optimum input pulley rotation speed.

(During constant torque ratio running) As shown in FIG. 31(A), the solenoid valves 84 and 85 controlled by the output of the electric control circuit are turned OFF.As a result, the oil pressure PI in the oil chamber 816 becomes line pressure. , the oil pressure P2 in the oil chamber 815 is also line pressure when the spool 812 is on the right side in the figure.The spool 812 is moved to the left in the figure because of the pressing force P by the spring 811.The spool 812 is moved to the left in the figure. Moved oil chamber 81
When 5 and drain boat 813 communicate with each other, the pressure in 22 is exhausted, so that the spool 812 is moved to the right in the figure by the oil pressure P in the oil chamber 816. When the spool 812 is moved to the right, the drain boat 813 is closed. In this case, by providing a flat slit 812b at the land edge between the drain boat 813 and the spool 812, the spool 812 can be moved in a more stable state as shown in FIG. 31(A).
It is possible to maintain an equilibrium point at an intermediate position as shown in FIG.

In this state, the oil passage 2 is closed, and the oil pressure of the hydraulic servo 313 of the input end pulley 31 is applied to the output pulley 3.
Due to the line pressure applied to the hydraulic servo 323 of No. 2, V
It becomes compressed via the belt 33 and is eventually balanced with the hydraulic pressure of the hydraulic servo 323. Actually, since there is oil leakage in the oil passage 2 as well, the input side boob IJ 31 is gradually expanded and the torque ratio T changes in the direction of increasing. Therefore, as shown in the third engineering drawing (A>), the spool 812
When the drain boat 814 is in equilibrium, the spool 812 is closed so that the oil passage 1 is slightly open.
A flat notch 812a is provided at the land edge of the oil passage 2 to compensate for oil leakage in the oil passage 2.

(During upshift) As shown in FIG. 31(B), the solenoid valve 85 is turned on by the output of the electric control circuit. As a result, the pressure in the oil chamber 816 is evacuated, so the spool 812 is moved to the left in the figure. As the spool 812 moves, the pressure in the oil chamber 815 is also evacuated from the drain boat 813. However, due to the action of the spring 811, the spool 812 is set at the left end in the diagram.

In this state, the line pressure of oil passage 1 is supplied to oil passage 2 via boat 81B, so the oil pressure of hydraulic servo 313 increases, input side boolean IJ 31 operates in the direction of closing, and torque ratio T decreases. do. Therefore, the solenoid valve 85 is turned ON.
The upshift is performed by reducing the desired torque ratio by controlling the time as necessary.

(During downshift) As shown in FIG. 31(C), the solenoid valve 84 is turned on by the output of the electric control circuit, and the oil chamber 815 is evacuated.

The spool 812 is moved to the right in the figure by the line pressure of the oil chamber 816, the oil passage 2 is communicated with the drain boat 814 and the pressure is discharged, and the input pulley 31 is operated in the expanding direction to increase the torque ratio. By controlling the ON time of the solenoid valve 84 in this way, the torque ratio is increased and a downshift is performed.

As described above, the output hydraulic pressure of the torque ratio control valve 81 is supplied to the hydraulic servo 313 of the input side (drive side) pulley 31, and the line pressure is introduced to the hydraulic servo 323 of the output side (driven side) pulley 32. If the oil pressure of the input side hydraulic servo 313 is P3 and the oil pressure of the output side hydraulic servo 323 is Po, then P, /Pi have characteristics as shown in the graph of FIG. 32 with respect to the torque ratio T. For example, when driving with throttle opening θ = 50% and torque ratio T = 1.5 (3 points in the figure), loosen the accelerator and θ = 30%.
In this case, if P and /Pi are maintained as they are, the transition will be to the point in the figure where the torque ratio T = 0.87, and conversely, if the torque ratio T = 1.5 is maintained, the input side The value of Pa/Pt is increased by the output of the torque ratio control mechanism 80 that controls the pulley and changed to the value at point C in the figure. In this way, P,
By controlling the value of /P as necessary, it is possible to set an arbitrary torque ratio corresponding to any load condition.

[Brief explanation of the drawing]

Fig. 1 is a schematic diagram of a V-belt type continuously variable transmission for vehicles, Fig. 2 is a diagram showing one embodiment of a hydraulic control circuit in the continuously variable transmission of the present invention, and Fig. 3 explains the operation of a manual valve. 4 is a diagram for explaining the operation of the detent valve and the throttle valve. FIG. 5 is a diagram for explaining the operation of the torque ratio valve. FIG. 6 is a diagram for explaining one embodiment of the present invention. 7 is a diagram showing the required line pressure characteristics of the hydraulic control circuit, FIG. 8 is a diagram showing the throttle pressure characteristics, and FIGS. 9, 10, and 11 are diagrams showing the characteristics of the present invention. FIG. 12 and FIG. 13 are diagrams for explaining the flow of processing in the electric control circuit, and FIG. 14 is a diagram for explaining the operation of the shift control mechanism. , FIG. 15 is a waveform diagram of the control pulse, and FIG. 16 is a waveform diagram of the control pulse.
Figure 17 shows the characteristics of the hydraulic pressure supplied to the input and output side hydraulic servos, Figure 17 shows the characteristics of the solenoid pressure, Figure 18 shows the characteristics of the output oil pressure of the shift control valve,
Figure 9 is a diagram for explaining shift shock control processing, Figure 20 is a diagram showing the best fuel consumption power line of the engine,
Figure 21 is a diagram showing the characteristics of engine output performance, Figure 22
Figure 23 shows the performance curve of the fluid transmission mechanism, Figure 23 shows the equal fuel consumption rate curve of the engine, Figure 24 shows the best fuel efficiency fluid coupling output curve, and Figure 25 shows the best fuel efficiency fluid coupling. Figures 26, 27, and 29 are diagrams showing the characteristics of the output rotation speed, and Figures 28 and 30 are diagrams for explaining the flow of processing in the electric control circuit.
FIG. 31 is a diagram for explaining the operation of the torque ratio control device; FIG. 32 is a diagram showing the relationship between the torque ratio and the pressure ratio of the input/output side hydraulic servo; FIG. 33 is a schematic diagram of a conventional continuously variable transmission. 30...V belt type continuously variable transmission, 31...Input side pulley, 32...Output side pulley, 33...V belt,
60... Hydraulic adjustment device, 80... Torque ratio control device, 311.321... Fixed flange, 312.322
...Movable flange, 313.323...Hydraulic servo. Figure 1 Applicant: Aisin AW Co., Ltd. Representative Patent Attorney Hiroki Shirai (5 others) Figure 4 (A) (Snake) (K9/Cm2) Figure 5 (Snake) Figure 5 Fig. 6 Fig. 9 Suittor r lock, Tee 12th ward Fig. 13 Pc or Pb (kg/cm') Fig. 15 Fig. 16 T71y, 8 (-ku Fig. 19 Fig. 22 D L / ! ;Otsu Figure 23 Figure 20 Figure 21 Engine rotation (rPm) Figure 24 Figure 25 Throttle r Noe a G4) Figure 26! Figure 29 Figure 30 Figure 27 Figure 28 t3 Torque ratio, T Figure 33

Claims (2)

[Claims] (1) Regulating the hydraulic pressure generated in the input-side pulley and output-side pulley that have a fixed flange, a movable flange, and a hydraulic servo that displaces the movable flange, a V-belt stretched between these pulleys, and a hydraulic power source. A hydraulic pressure adjustment device; and a torque ratio control device that selectively supplies the hydraulic pressure regulated in the hydraulic pressure adjustment device to the hydraulic servo to make the effective diameters of the input pulley and the output pulley variable; The hydraulic pressure regulated in the hydraulic pressure adjustment device is controlled to the minimum necessary pressure using the engine output torque and the torque ratio between the input and output pulleys as parameters, and the pressure is increased at and around the maximum torque ratio. A speed change control device for a V-belt continuously variable transmission for vehicles. (2) The pressure is increased when the throttle opening corresponding to the output torque of the engine is equal to or higher than a predetermined value. Machine speed control device.