JPH023772A - Speed change control device for v-belt type continuously variable transmission in vehicle - Google Patents

Abstract

PURPOSE:To prevent a speed change shock, when a shift N-D is performed, by providing in a shift control mechanism a solenoid valve for controlling a supply oil pressure to a friction engaging unit engaging or releasing a predetermined rotary element of a planetary gear mechanism. CONSTITUTION:When shifts N-D and N-R are performed, a riseup of supply oil pressure Pb or Pc to a hydraulic servo 48 or 49 from a shift control valve 71 of a shift control mechanism 70 is controlled so as to obtain a predetermined curve by duty-driving a solenoid valve 74. Thus because the supply oil pressure to the hydraulic servos 48, 49 in friction engaging units 42, 45 of a forward- reverse switching mechanism 40 is controlled by a duty control of the solenoid valve 74, a speed change shock can be prevented when the shift is performed from an N-range to a D-range or to an R-range.

Description

Detailed Description of the Invention [Industrial Field of Application] The present invention provides a shift control device for a V-belt continuously variable transmission for a vehicle to prevent sluggishness when shifting from N range to D range or R range. Regarding.

[Conventional technology]

The V-belt type continuously variable transmission uses an endless metal V-belt invented by Pan d'Orne wrapped around two pulleys. It is characterized by the fact that the mold blocks are arranged without interruption, and power is transmitted between the sides of the boob and the side surfaces of the ■-shaped blocks and by the pressing force between the ■-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 that control power transmission according to the torque ratio, and forward/reverse switching mechanisms that prevent sluggishness when shifting from N range to D or R range, but no effective proposals have been made yet. .

However, as a prior invention, British Patent No. 2.058.250
No. (Borgoo Warner Limited, September 7, 1972)
(Application filed on April 8, 1981, Publication issued on April 8, 1982) is known. This consists of a torque converter connected to the output shaft of the engine, a continuously variable transmission mechanism with a fixed pulley and a movable pulley on the input and output shafts, and a V-belt wrapped between the input and output shafts, and a planetary gear. It has a forward/reverse switching mechanism using a unit and a reduction gear mechanism connected to a differential gear device, and a hydraulic control device moves the movable boob to change the torque ratio, and the planetary gear unit Forward and reverse movement is controlled by engaging or disengaging predetermined rotating elements.

[Problem to be solved by the invention]

By the way, in the torque ratio change and forward/reverse switching control described above, the difference between this type of continuously variable transmission and a general automatic transmission is that the continuously variable transmission transmits power between the metal pulley and the metal block. Since this operation is performed with low frictional resistance, high hydraulic pressure is required, and even when the engine is idling (throttle opening 0), V during engine braking is
This requires relatively high hydraulic pressure to prevent the belt from slipping.

On the other hand, in the forward/reverse switching mechanism, a clutch or brake consisting of a large number of friction plates engages a predetermined rotating element of the planetary gear unit, similar to a general automatic transmission.

Therefore, in the prior patent, the high working oil pressure for changing the torque ratio is applied to the clutch of the forward/reverse switching mechanism.
Because it is used for brake hydraulic pressure, the neutral range (
When shifting from the N range to the driving range (D range or R range), the clutch or brake is rapidly engaged, causing a shift shock.

The present invention solves the above-mentioned problem, and is intended to change the range from the neutral range (N range) to the running range (D range or R range).
An object of the present invention is to provide a shift control device for a V-belt type continuously variable transmission for a vehicle, which can prevent shift shock when shifting to a vehicle.

[Means to solve the problem]

To this end, the present invention provides a fluid transmission device (21), a forward/reverse switching mechanism (40), and a V-belt continuously variable transmission mechanism (30).
In the V-belt type continuously variable transmission for a vehicle, the forward/reverse switching mechanism (40) includes a friction engagement device (42, 45) that engages or releases a predetermined rotating element (43, 46) of the planetary gear mechanism. and a hydraulic power source (50) for the frictional engagement device.
and a shift control mechanism (70) that supplies hydraulic pressure from the friction engagement device (4).
2, 45) is provided with a solenoid valve (74) for controlling the oil pressure supplied to the hydraulic pressure.

[Action and effect of the invention]

In the present invention, when relieving locking stiffness during N-D shift and N-R shift, the hydraulic servo 48
Alternatively, the rise of the hydraulic pressure P or PC supplied to the hydraulic servo 49 is controlled, for example, as shown in the hydraulic characteristic curve shown in FIG. 16, and the engagement of the multi-disc clutch 45 or the multi-disc brake 42 between AC is completed in the diagram. urge The relationship between the duty (%) of the solenoid valve 74 for controlling the hydraulic pressure supplied to the hydraulic servo 48 or 49 and the solenoid pressure P generated in the oil chamber 713 by the operation of the solenoid valve 74 is shown in FIG. The solenoid pressure shown in FIG. 22 is amplified by the shift control valve 71 and
Hydraulic pressure P supplied to the hydraulic servo 48 or 49 shown in FIG.
Or you can get a PC.

According to the present invention, in order to control the hydraulic pressure supplied to the friction engagement device of the forward/reverse switching mechanism using the solenoid valve, the hydraulic pressure supplied to the friction engagement device of the forward/reverse switching mechanism is controlled from the neutral range (N range) to the travel range (D range or R range).
It is possible to prevent shift shock when shifting to range).

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. The continuously variable transmission device consists of a planetary gear transmission 40 for forward/reverse switching connected to the output shaft 26 of the continuously variable transmission 1130, and a principle gear mechanism 23 connected to the output shaft 47 of the planetary gear transmission @40. ing.

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 V-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 30, and a fixed flange 321 connected to the output shaft 26 of the continuously variable transmission 30.
and a movable flange 322 provided so as to face and 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 displacement IL of the movable flanges 312 and 322 of the input pulley 31 and output pulley 32 is 0 to 12 to 1,
~It4 (0<12<13<14), so that the torque ratio T between the input shaft 214 and the output shaft 26 is t, xlz -wt, ~j4 (t+ <tz
<t3 <t4) Continuously variable speed is performed. In this embodiment, the pressure receiving area of the input side hydraulic servo 313 is approximately twice as large as the pressure receiving area of the output side hydraulic servo 323, and the hydraulic pressure applied to the hydraulic servo 313 is equal to or equal to the hydraulic pressure applied to the hydraulic servo 323. Even if the input side movable flange 312 is small, the movable flange 312 on the input side is formed to obtain a larger driving force than the movable flange 322 on the output side. This increase in the pressure receiving area of the hydraulic servo 313 is achieved by increasing the diameter of the hydraulic servo or by employing a piston having a double pressure receiving area in the hydraulic servo.

The planetary gear transmission 41140 for forward/reverse switching is sun gear 4.
1, a ring gear 43, a sun gear 41, a double planetary gear 44 that meshes with the ring gear 43, and a carrier 46 that rotatably supports the double planetary gear 44. The sun gear 41 is connected to the output shaft 26 of the continuously variable transmission. , the carrier 46 is connected to an output shaft 47 of a planetary gear transmission 40 for forward/reverse switching. 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 used to compensate for the fact that the speed change range obtained by the belt-type continuously variable transmission a, 30 is lower than the speed change range achieved by a normal vehicle transmission, and has a reduction ratio of 1.45, for example. The torque is increased by decelerating the engine. 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 belt-type continuously variable transmission shown in FIG.

The hydraulic control circuit includes a hydraulic power source 5o, a hydraulic adjustment device 60, and an N-
It consists of a shift control mechanism 70 that controls the impact at tII8I when D and N-R226, and a torque ratio control device 8°.

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

The hydraulic adjustment device 6o includes a manual valve 62 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 side. A torque ratio valve 66 that operates in conjunction with the movable flange 322 of the pulley 32 and supplies line pressure to the detent valve 64 according to its changing position, and discharges pressure from the output hydraulic pressure feedback oil passage 9 provided in the throttle valve 65, and the hydraulic pressure 'rA50. It is composed of a regulator valve 61 that applies pressure to the hydraulic pressure supplied from the servo and supplies it to the hydraulic servo 323 as line pressure.

The manual valve 62 has a spool 621 in positions P, R, and N, as shown in FIG. 3, corresponding to shift positions P, R, N, D, and L of a shift lever provided at the driver's seat.

It is set at each position D and L, and connects the oil passage 1 to which line pressure is supplied and the output oil passages 3 to 5 as shown in Table (2).

In the table (■), O 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 valve 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 1. The oil discharged from port 614 is supplied to the fluid compling, oil cooler, and lubrication points via oil path I2.

The detent valve 64 includes a spool 641 that moves in conjunction with the throttle opening θ of the butterfly valve of the carburetor 102 as shown in FIG.
≦θ, as shown in FIG.
When it is 1.0%, 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 includes a spool 651 which is arranged in series with a detent valve 641 via a spring 645 and has a spring 652 on its back. Due to the action of the spool 651 that moves in accordance with the fluctuation of the degree θ, the boat 65 connected to the oil passage l
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. It receives a hydraulic feedbank.

The torque ratio valve G6 includes a spool 6 linked to the movable flange 322 of the output pulley 32 via a connecting rod.
62, the amount of movement of the movable flange 322! ,≦L
≦14 (Torque ratio T is 12≧T≧1+), the fifth
As shown in Figure (A), the spool 662 is located on the left side of the figure, and closes the input port 64 connected to the feedback oil passage 9 of the output oil pressure provided in the throttle valve 65, and supplies the output oil to the detent valve 64. Channel 6 is communicated with drain boat 665 to evacuate pressure. The movement IL of the movable flange 322 is smaller than the first set value &, 7!2 ≦L
<12 (t3≧T>t), the spool 662 is located in the middle as shown in FIG. is depressurized.

When the movement ill of the movable flange 322 is smaller than the second set value 12 and O≦L≦12 (t4≧T>t,), the spool 662 is positioned on the right side as shown in FIG. 5(C). A boat 663 and oil passage 6 connected to oil passage 1
are in communication 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 placed behind it on one end and receiving line pressure from an oil chamber 713 provided at the other end, and an oil chamber 713.
Orifice 7 provided in the oil passage l that supplies heline pressure
2, a plunger limiting valve 73 installed between the orifice 72 and the oil chamber 713, and a solenoid valve 74 that adjusts the oil pressure in the oil chamber 713 under the control of an electric control circuit to be 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.

Drain when solenoid valve 74 is off)
741 is closed, and the spool 712 is located on the right side in the figure with line pressure supplied to the oil chamber 713, and connects the oil passages 3 and 4 to the oil passages 13 and 14, respectively, and connects the multi-disc brake 42. Alternatively, the multi-disc clutch 45 is engaged.

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 engaged.

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 having a spring 811 on its back, oil chambers 815 and 816 at both ends to which line pressure is supplied from the oil passage 1 through orifices 82 and 83, and line pressure. Oil road 1
An input 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 path 2. An oil chamber 819 provided, a drain boat 814 that evacuates the oil chamber 819 according to the movement of the spool 812, and a drain boat 813 that evacuates the oil chamber 815 according to the movement of the spool 812.
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 and torque ratio control device 280 of the shift control mechanism 70 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, RSN, D, or L; 902 is a rotational speed sensor that detects the rotational speed of the input side boolean IJ31; 903 is a vehicle speed sensor; 904 is a A throttle sensor 905 detects the throttle opening of the carburetor or the amount of depression of the accelerator pedal, and 905 is a rotation speed sensor 902
906 Speed detection processing circuit that converts the output of
is a vehicle speed detection 1 that converts the output of the vehicle speed sensor 903 into voltage.
A circuit 907 is a throttle opening detection processing circuit that converts the output of the throttle sensor 904 into voltage, 908 to 911
912 is the central processing unit (CPU), 913 is the solenoid valve 74.84.
Read-on memory (ROM), 914, which stores programs to control the 85 and data necessary for control.
915 is a random access memory (RAM) that temporarily stores input data and parameters necessary for control.
916 is a clock, 916 is an output interface, and 917 is a solenoid output driver, which converts the output of the output interface 916 into operating outputs of the upshift solenoid valve 85, the downshift solenoid valve 84, and the shift control solenoid valve 74. Input interfaces 908 to 911, CPU 912, ROM 913
, RAM 914, and output interface 916 are connected through a data bus 918 and an address bus 919.

Next, torque ratio valve 66, detent valve 64, throttle valve 65, manual valve 62, and regulator valve 6.
1 will be described.

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 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. The hydraulic pressure is defined by the hydraulic pressure at which the hydraulic servo can transmit torque without causing the V-belt 33 to slip.

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 have a line pressure as shown by a large broken line, and even when the throttle opening θ=0 during engine braking, it is desirable to have a higher line pressure characteristic as shown by a dashed line.

In this embodiment, the line pressure, which is the output of the regulator valve 61, is controlled by the manual valve 6 by the hydraulic pressure adjusting device 6o.
2 shift position 1j (L, D, N, R.

P) is adjusted as follows by changing the throttle opening θ and the torque ratio between the input and output shafts.

(D position 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 passages 4 and 5 are exhausted. At this time, the shift control mechanism 70 , when the shift control solenoid valve 74 is in the OFF state, 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 boat 663 connected to the drain boat 665 is closed, and the oil passage 6 is communicated with a drain boat 665 to discharge pressure.

As a result, no detent pressure is generated in the oil passage 7, regardless of the slow/)le opening degree θ. Also, the throttle valve 65
is the boat 66 of the torque ratio valve 66 connected to the oil line 9.
4 is closed, and the spool 651 receives feedback pressure from the land 657 in addition to the land 656.
A throttle pressure having the characteristics shown in FIG. 8(c) with respect to the throttle opening θ is outputted to the regulator pulp plunger 612 of the regulating valve 61 via the oil passage 8. As a result, the line pressure output from the regulator 61 becomes as shown in the area (f) of FIG. 9 and the area (e) of FIG. 10.

■ When the torque ratio T is t, <"r≦t.

As shown in FIG. 5(B), the torque ratio valve 66 is connected to the boat 6.
63 is closed, allowing the oil passage 9 and the drain boat 666 to communicate with each other. The oil passage 6 is also depressurized through the boat 665. Therefore, no detent pressure is generated, and the throttle pressure increases by the amount that the pressure in the oil passage 9 is exhausted and the feedback pressure is no longer applied to the land 656 of the spool 651, and is expressed by the characteristic curve shown in FIG. 8(d). Ru. The line pressure at this time has the characteristics shown in (L) 1 in FIG. 9 and (G) in FIG.

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

As shown in FIG. 5(C), the oil passage 9 is connected to a drain boat 666.
Therefore, the throttle pressure 7) is the same as above (■).
This is shown in figure (d). However, since the bow) 663 is opened and the oil passage 1 and the oil passage 6 are communicated with each other, the throttle opening θ becomes 0.
≦θ≦01% and while the spool 641 of the detent valve 64 is on the left side of the figure as shown in FIG. is discharged from the manual valve 62 via the oil passage 5, but when the throttle opening θ is between 01% and 05100%, the spool 641 moves to the right 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 increases as shown in Fig. 9 (
As shown in area (w) and (s) in Figure 10, θ=θ2%
The characteristic changes stepwise.

(L position) The oil passage 5 communicates with the oil passage 1 in the manual valve 62 .

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

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

When the throttle opening degree θ is 050508%, the oil passage 5 and the oil passage 7 communicate with each other in the detent valve 64, and detent pressure is generated to push up the throttle plunger and generate high line pressure. Throttle opening θ is 01% <0
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, <'l'≦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, <7'≦t4.

The oil passage 6 and the oil passage 1 are communicated with each other by the torque ratio valve 66, and the pressure of the oil passage 9 is discharged from the drain boat 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 inputs the detent pressure and the same throttle pressure as in 7) above. The regulator valve 61 outputs the line pressure shown in FIG.

(R position) At the manual valve 62t, oil passages 4 and 5 communicate with oil passage 1, and oil passage 3 is depressurized. At this time, in the shift control mechanism 70, the shift control solenoid 74 is
When line pressure is being supplied to the oil chamber 713 in the FF 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 t, <T≦t, as shown in the characteristic curve (S) in Fig. 10, is set at a throttle opening of θ1% or less. The reason why the line pressure is set low is when the line pressure is set high under operating conditions such as idling where the throttle opening θ is small and the pump discharge amount is low, and 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, as it becomes difficult to maintain the line pressure, and furthermore, the oil temperature further increases due to a decrease in the amount of oil supplied to the oil cooler, which is likely to cause trouble.

In addition, when the manual valve 62 is shifted to the G and R shift positions, the characteristic curves (C) and (L) in FIG.
As shown in , when the torque ratio T is 1. The line pressure was set high under operating conditions in the range of ≦T≦t2 and the throttle opening θ is 0.1% or less because relatively high oil pressure is required even at low throttle opening during engine braking. by.

In this way, as shown in FIG. 9, by bringing the line pressure close to the minimum required oil pressure shown in FIG. 7, the power loss caused by the pump 52 can be reduced, thereby improving fuel efficiency.

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 large force pulley 31 is brought into a neutral state. When the shift lever changes from the P position or N position to the R position, and from the N position to the D position, a shift shock is applied to alleviate the shift shock associated with the N-R shift and N-D shift, respectively. Control processing is performed (steps 940 and 950).

This shift shock control processing is a feature of the present invention and will be described 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
S, , S, , SS2, spring 71 in order from the left side of the figure
1's elastic force is Fs+, and the oil pressure of the oil chamber 713 is P, the hydraulic servo 4 of the multi-disc clutch 45 that is engaged when moving forward
The hydraulic pressure PC supplied to 9 and the hydraulic pressure P supplied to the hydraulic servo 48 of the multi-disc brake 42 that is engaged during reversing are the 0th equation and the 0th equation, respectively, which are hydraulic balance equations of the shift control valve 71.
From the equation, it is given as follows.

When moving forward P, XS, = Pcx3. +F! , ■S!
sz When moving backward P, xS, = Pb
If the pressure receiving area of 1 is 53, and the elastic force of the spring 732 placed behind the valve body 731 is F, then the pressure lifting valve 73 operates at the highest pressure pli+sit of P according to the hydraulic balance equation No. 1. do.

Pb/1m1tXsz =Fsz
■P 1 imi t= F sz/ S 3 At this time, PC and P are the highest pressure P according to the 0th formula and the 2nd formula.
c11m1t, Pb/1m1t are limited.

When moving forward St St When moving backward L Pb j! imonit=P
It fitS+ St The solenoid valve 74 has a duty (%
), a solenoid pressure P is generated in the oil chamber 713.

Duty (%) = This duty control has a pulse width of L"-nM" (n-1, 2,
3,...), and the pulse gradually becomes smaller during the shift control solenoid valve 7 shown in FIG.
This is done by adding to 4. By duty-controlling the shift control solenoid valve 74 in this manner, a hydraulic pressure P adjusted in accordance with 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 Pb or P6 supplied to 9 is rolled as shown in the oil pressure characteristic curve shown in FIG.
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 Ml. In step 941, it is determined whether or not PLUG is on during shock control processing, and FLO
If G is on, shift shock control processing is in progress and the process proceeds to step 946; if PLUG is not on, RA is turned on to start shift shock control processing.
By comparing the shift lever position stored in M914 with the current shift lever position, it is determined whether the shift lever has changed from the P position or 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, the corresponding parameters K", L", M" are changed in steps 944 and 945.
PLUG that indicates that the parameter K is set to 0 and shock control processing is performed.
is turned on (step 955), and if no change has occurred, the process returns and no shift control processing is performed.

In step 946, it is determined whether the parameter K that determines the end of 0 in the - period is greater than O. If K is not greater than O, K is set to 0-1, L is set to L",
Set Ll as L"-Ml (step 947), and in step 948 determine whether L is less than or equal to O. If L is less than θ, proceed to step 951; if L is less than or equal to O, all It is assumed that the shock control process has been completed and the PLUG is turned off.
When the parameter K, which determines the end of the ON time, is larger 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 0 (step 950). 951). When L is 0, a command to turn off the solenoid valve 74 is issued and the process proceeds to step 95.
2) If L is not O, issue an on command and step 9
53) After that, set L-1 to L and return.

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

The explanation will be returned to FIG. 12.

Next to the N-D shift shock control section F1950, the actual input pulley rotation speed N is read by the input pulley rotation speed sensor 902 (step 923).
Next, it is determined whether the throttle opening degree θ read in step 921 is 0 or not (step 924), and when θ≠0, the input pulley target rotation speed N0 is set to the input pulley rotation speed for the best fuel economy. Execute subroutine 960;
When θ=0 and the throttle is fully closed, 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), and the shift lever is set to the D position. When the shift lever is set to the L position, an engine brake processing subroutine 970 for the D position is executed, and when the shift lever is set to the L position, an engine brake processing subroutine 980 for the L position is executed, and the target rotation speed of the input pulley is set. N9
Set to appropriate values for each.

A subroutine 960 for setting the input pulley target rotation speed N0 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 5-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 found. If the fuel consumption rate is Q (g/ps-h) and the horsepower is P (pS), 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, when the engine 100 and the fluid coupling 21, which is a fluid transmission mechanism, are combined as in this embodiment, the engine output performance curve at the slow opening angle θ shown in FIG. Then, from the fluid coupling performance curve shown in FIG. 22 and the engine equivalent fuel efficiency rate curve shown in FIG. 23, the best fuel efficiency fluid coupling output line is found on the fluid coupling output performance curve shown in FIG. 24. I can do it. FIG. 25 shows the best fuel efficiency fluid coupling output line shown in FIG. 24 replaced by 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 N0 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 Step 961), select the data address, read out the best fuel efficiency input end rotation speed N from the set address, step 962), and read out the best fuel efficiency input end boolean corresponding to the read throttle opening θ. The data of the rotational speed N is set to the input end pulley target rotational speed N° (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.
(971), calculate the acceleration α at that point (972), and then calculate the acceleration α, which is an appropriate acceleration A for the vehicle speed.
It is determined whether or not it is (973). When the acceleration α is larger than the acceleration A, the input pulley target rotation speed I" should be set to a value larger than the current input pulley rotation speed N in order to downshift (974), and the acceleration α is not larger than the acceleration A. In some cases, the input pulley target rotation speed N
0 is set to the best fuel efficiency input end pulley rotation speed N corresponding to the throttle opening 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 transmisosilane, the final reduction ratio of the vehicle, the tire radius, etc. Next, it is determined whether the current torque ratio T is smaller than the torque ratio T0 that provides safe and appropriate engine braking for the vehicle speed.
When is smaller than the torque ratio T1, the input pulley target rotation speed N0 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 T'', the input pulley target rotation speed N'' is set to the current input pulley rotation speed N (985), and the process returns. The torque ratio T11 at which safe and appropriate engine braking can be obtained 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, the actual input pulley rotation speed N and the best fuel efficiency input pulley rotation speed N'' are compared in FIG. Step 927), N (If No, issue a command to operate the downshift solenoid valve 84. Step 928)
, N>N'', an operation command for the amplifier shift solenoid valve 85 is issued (step 929), and 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 change in the gear ratio between the human output pulleys by comparing the rotation speed of the input pulley for the best fuel consumption obtained in FIG. This is done by operating two solenoid valves 84 and 85 provided in the control device 80, and the actual input side pulley rotation speed is made to match the input side pulley rotation speed for the best fuel efficiency.

(During constant torque ratio running) As shown in FIG. 31(A), solenoid valves 84 and 85 controlled by the output of the electric control circuit are turned off. As a result, the oil pressure P1 in the oil chamber 816 becomes the line pressure, and the oil pressure P2 in the oil chamber 815 also becomes the 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 P3 exerted by the spring 811. The spool 812 is moved to the left and the oil chamber 81
5 and the drain port 813 communicate with each other, pressure P2 is exhausted, and the spool 812 is pushed to the right in the figure by the oil pressure P in the oil chamber 816. When spool 812 is moved to the right, drain port 813 is closed. In this case, by providing a flat notch 812b at the land edge of the drain port 813 and the spool 812, the spool 812 can be held in a more stable state as shown in FIG.
), it is possible to maintain an equilibrium point at an intermediate position.

In this state, the oil passage 2 is closed, and the oil pressure of the hydraulic servo 313 of the input 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, the input pulley 31 is gradually expanded and the torque ratio T changes in the direction of increasing (as shown in FIG. 31(A), In the position where spool 812 is balanced, drain boat 814 is closed;
A flat notch 812a is provided at the land edge of the oil passage 1 with the spool 812 so that the oil passage 1 is in a slightly open state 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 818, so the oil pressure of hydraulic servo 313 increases, input pulley 31 operates in the direction of closing, and torque ratio T decreases. do. Therefore, by controlling the N time of the solenoid valve 85 as necessary, the torque ratio is reduced by a desired amount and the amplifier shift is performed.

(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 P8, and the oil pressure of the output side hydraulic servo 323 is Po, then P0/P has a characteristic with respect to the torque ratio T as shown in the graph of FIG. For example, when driving with throttle opening θ = 50% and torque ratio T = 1.5 (point a in the figure), loosen the accelerator and θ = 30%.
In this case, when P and /Pi are maintained as they are, the torque ratio shifts to the point in the figure of T-0, 87, and conversely, when the torque ratio T=i5 is maintained, the input pulley P, /P by the output of torque ratio controller 1J180 that controls
The value of i is increased and changed to the value at point C in the figure. By controlling the value of P0/Pe as necessary in this way, 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, and FIG. 6 is a diagram for explaining one embodiment of the present invention. FIG. 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 configuration of the electric control circuit. A diagram showing line pressure characteristics obtained by the control device, FIGS. 12 and 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. Figure 15 is a waveform diagram of control pulses, Figure 16 is a waveform diagram of control pulses.
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, and 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 consumption fluid compression 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; Figures 26, 27, and 29 are diagrams explaining the flow of processing in the electric control circuit; Figures 28 and 30 are diagrams explaining control setting data. FIG. 31 is a diagram for explaining the operation of the torque ratio control device, and FIG. 32 is a diagram showing the relationship between the torque ratio and the pressure ratio of the input/output side hydraulic servo. 21...Fluid transmission device, 30...Continuously variable transmission mechanism, 3
3...■ Belt, 40... Forward/forward switching mechanism, 42.
45...Friction engagement device, 70...Shift control mechanism,
74...Solenoid valve. Applicant: Aisin AW Co., Ltd. Representative Patent Attorney Hiroki Shirai (5 others) Fig. 4 (A) (B) Fig. 6 Fig. 6 (K9/Cm2)] 81 Fig. 9 Between the throttles is O Roof 13 Fig. 12 Fig. 15 Fig. 16/'lνδ Medical darkness Pc Mata 1yo Pb (kg/crn”) Fig. 19 s Jun-Hachi ni 7 history Fig. 20 21 Figure Ensonoro Tenho (rpm) Figure 22 Yu'',! Figure 23 Figure 26 Figure 24 Figure 25 Figure 27 Figure 28 Figure 29 Figure 30

Claims (3)

[Claims] (1) In a vehicle V-belt continuously variable transmission having a fluid transmission device, a forward/reverse switching mechanism, and a V-belt continuously variable transmission mechanism, the forward/reverse switching mechanism engages or disengages a predetermined rotating element of a planetary gear unit. the friction engagement device, and a shift control mechanism that supplies hydraulic pressure from a hydraulic source to the friction engagement device, and the shift control mechanism includes a solenoid for controlling the hydraulic pressure supplied to the friction engagement device. A speed change control device for a vehicle V-belt type continuously variable transmission characterized by being provided with a valve. (2) The shift control mechanism is configured to adjust the hydraulic pressure supplied from the hydraulic pressure source and adjusted by the hydraulic pressure adjustment device to a predetermined pressure required for the hydraulic servo of the V-belt type continuously variable transmission mechanism by the solenoid valve. A shift control device for a V-belt type continuously variable transmission for a vehicle according to claim 1, further comprising a shift control valve that adjusts the hydraulic pressure supplied to the frictional engagement device according to a solenoid pressure. . (3) The shift control valve includes a first oil chamber to which oil pressure is supplied to the frictional engagement device, a second oil chamber to which the solenoid pressure is supplied, and a first oil chamber and a second oil chamber. selectively connecting a first oil passage to which the hydraulic pressure supplied to the chamber acts oppositely to supply the hydraulic pressure regulated by the hydraulic pressure adjustment device and a second oil passage connected to the frictional engagement device; The second oil chamber communicates with the first oil passage through an orifice, and is provided with a pressure limiting valve that maintains the oil pressure of the second oil chamber below a predetermined value. A speed change control device for a V-belt type continuously variable transmission for a vehicle according to claim 2.