JPS63176746A - Speed change controller for v-belt type continuously variable transmission for vehicle - Google Patents

Abstract

PURPOSE:To improve feeling of engine brake by setting the speed change target values having different characteristics through plural means and selecting/setting one of them corresponding to operating condition. CONSTITUTION:Rotation N'' of an input side pulley 30 having best specific fuel consumption corresponding to a throttle opening is compared with an actual rotation N, and a speed change command is produced. When a throttle is fully closed, it is judged whether a shift lever is set at L-position and an engine brake is operated as required. At position D, a vehicle speed sensor 903 reads in a vehicle speed V and calculates an acceleration (alpha) at that time point so as to judge whether it is a proper acceleration A, and if alpha>A, N'' is set to a value higher than N and the engine brake is operated. At position L, the current torque ratio is compared with a torque ratio with which proper engine brake is obtained for the vehicle speed V. Consequently, the engine brake operation can be switched to the best state.

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]

■Belt-type continuously variable transmissions can be used as automatic transmissions for automobiles and other vehicles in combination with planetary gear transmissions for forward and reverse switching. Figure 30 is from Japanese Patent Application Laid-Open No. 54-1579.
A conventional continuously variable transmission proposed in Publication No. 30 is shown,
A fixed pulley b and a movable pulley C are provided on the human power shaft a.
Further, a fixed pulley e and a movable pulley f are provided on the output shaft d, and a belt g is stretched between the input shaft a and the output shaft d, and fluid from the pump j is connected to the oil path and the oil path i. The movable pulley c is supplied to and discharged from the valve via N.
, f are moved.

On one end of the spool m of the valve l, a fluid pressure proportional to the rotation speed of the input shaft a is applied by a pitot tube n, and on the other end of the spool m, the rotation of a cam p that is linked to the movement of the throttle pedal is applied. The pressure due to the movement is acting through lever q and spring r. Furthermore, fluid pressure proportional to the rotational speed of the human power shaft a is applied to one end of the spool S of the valve by a pitot tube n, and the other end of the spool S is connected to the movable pulley C of the input shaft a. The pressure of the detection iron t that is displaced by the lever U and the spring V acts on the lever U and the spring V.

In the above configuration, the oil passage is controlled by the balance between the load of the spring r that changes due to the rotation of the cam p linked to the movement of the throttle pedal, and the fluid pressure generated by the pitot tube n in accordance with the rotation speed of the human power shaft a. The hydraulic pressure generated in this oil path is controlled by the hydraulic pressure generated in the cam p, the lever q, the spring r, the pitot tube n, etc. It is subject to change.

[Problem that the invention seeks to solve]

However, the speed change control in the conventional continuously variable transmission described above is performed by moving the movable boob IJ (IJ) using hydraulic pressure supplied from the valve l, but in vehicles equipped with the continuously variable transmission,
When changing the speed change characteristics based on the driver's will (for example, the position of the shift lever), it is necessary to change the output hydraulic characteristics of the cam p, lever q, spring r, pitot tube n, etc. There are problems in that it is not easy to change the speed change characteristics, and it is also not easy to change the speed change characteristics due to differences in driving conditions.

Furthermore, although the above-mentioned gear shifting is performed in accordance with the best fuel economy curve of the engine, in the coasting state with the accelerator pedal released, the continuously variable transmission shifts toward the lowest torque ratio (overdrive side), and as a result, There is a problem in that the vehicle coasts with almost no feeling of deceleration, and further accelerates when exiting the vehicle. On the other hand, if the engine brake is set so that it is always applied when the accelerator pedal is released, there is a problem in that even when the driver wants to coast, the driver feels a strong sense of deceleration and worsens the driving feeling.

The present invention solves the above-mentioned problems, and is a belt-type non-vehicle vehicle that can easily change the speed change characteristics and provide speed change performance that is suited to the vehicle condition and the driver's will. An object of the present invention is to provide a speed change control device for a step-change transmission.

[Means for solving problems]

For this purpose, the speed change control device of the V-belt type continuously variable transmission for vehicles of the present invention is attached to the input shaft and the output shaft, respectively.
For vehicles, which consists of an input pulley and an output pulley with variable effective diameters, and a drive band stretched between these pulleys, and controls the torque ratio between the input and output shafts by adjusting the effective diameter of the pulleys. In the V-belt continuously variable transmission, means for detecting vehicle operating conditions such as throttle opening, vehicle speed, input pulley rotation speed, and means for setting a gear shift target value in response to a signal from the detection means; An electric control circuit that compares these speed change target values and the detected value and outputs a signal based on the comparison result, and a torque ratio control means that adjusts a torque ratio based on an output signal of the electric control circuit, The means for setting the target value consists of a plurality of means for setting shift target values having different characteristics, and one of the means is selected from the plurality of means according to the driving condition of the vehicle to set the shift value +S value. This is a characteristic feature.

[Action and effect of the invention]

In the present invention, for example, as shown in FIG. 27, the control of the torque ratio control device 80 involves comparing the input pulley rotation speed for the best fuel consumption obtained in FIG. 17 with the actual input pulley rotation speed. Two solenoid valves 84 provided in the torque ratio control device 80 increase the speed ratio between the input and output pulleys.
and 85, and the actual input pulley rotation speed is made to match the input pulley rotation speed for the best fuel efficiency.

That is, as shown in FIG. 18, the input pulley rotation speed sensor 902 reads the actual input pulley rotation speed N (923), and then it is determined whether the throttle opening θ is 0 or not (924). ), when θ≠0, set the best fuel efficiency input pulley rotation speed N" corresponding to the throttle opening θ shown in FIG. A comparison is made with the rotational speed N1 (927) and a shift command is issued, but when 0 = 0 and the throttle is fully closed, the shift lever is set to the D position or to the L position to judge the necessity of engine braking. It is determined whether the settings have been made (926), and engine brake processing 970 or 980 is performed as necessary.

As shown in FIG. 22, the engine brake processing 970 at position D reads the vehicle speed V using the vehicle speed sensor 903 (9
71), calculate the acceleration α at that point 972),
Next, it is determined whether the acceleration α is an appropriate acceleration A for the vehicle speed (973). When α>A, to perform downshift control 974, return after setting No to a value larger than N, and when α≦A, N1 is the best fuel economy input side pulley rotation speed N9 corresponding to throttle opening 0. After setting (975), the process returns.

In the engine brake processing 980 for the L position, as shown in FIG. 23, after reading the vehicle speed ■ (981), the torque ratio T is calculated from the vehicle speed ■ and the input pulley rotation speed N (982), and then the current It is determined whether the torque ratio T of is larger than the torque ratio T1 that provides safe and appropriate engine braking for the vehicle speed ■ (983),
＜T'', change N to N'' so that a downshift is performed.
A larger value is set (984), and when T≧T9, No is set to 1, which is equal to N, and the process returns (985).

According to the present invention, the effectiveness of engine braking can be switched by selecting the driving stroke, and engine braking with good feeling and performance can be obtained.
Optimal characteristics such as fuel efficiency and power performance can be obtained even in different vehicles.

〔Example〕

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

FIG. 1 is a schematic diagram of an automobile transmission using a continuously variable transmission.

100 is an engine, 102 is a carburetor, and 20 is a transmission device provided between the engine 100 and the drive side axle, and is connected to a hydraulic fluid coupling 21 and a differential gear 22 connected to the output side 101 of the engine. It is constituted by a continuously variable transmission including a reduction gear mechanism 23, a V-belt continuously variable transmission 30, and an a-star gear transmission 40 for forward/reverse switching.

The fluid coupling 21 is well known and consists of a pump impeller 211 and a turbine runner 212 coupled to a fluid coupling output shaft 214. Note that other fluid torque converters or mechanical clutches may be used instead of the fluid coupling.

■The belt type continuously variable transmission 30 has a fixed flange 311 connected to the fluid coupling output shaft 214, which is the human power shaft of the continuously variable transmission [30], and forms a V-shaped space facing the fixed flange 311. An input side boo U 31 consisting of a movable flange 312 provided to
A fixed flange 321 connected to the intermediate shaft 26 that is the output shaft of R30, a movable flange 322 provided to face the fixed flange 321 to form a V-shaped space, and a hydraulic servo that drives the movable flange 322. 323, and a V-belt 33 which is a driving 0J band and connects the input pulley 31 and the output pulley 32.

The displacement of the movable flanges 312 and 322 of the input pulley 31 and output pulley 32 is 0 to j! 2~l
, ~1.4 (Q < 12 < (23 < ea
), and as a result, the torque ratio T between the input shaft 214 and the output shaft 26 is L1~L2~L3~(4(mu,<1.<
13<1. ).Continuously variable speed is provided. 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 oil servo 313 is equal to the hydraulic pressure applied to the hydraulic servo 323. Even in the case where 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 southeast transmission 40 for forward and reverse switching includes a sun gear 4 connected to an intermediate shaft 26 that is the output shaft of the continuously variable transmission 30.
1. A ring gear 43 engaged with the transmission case 400 via a multi-plate brake 42, a double planetary gear 44 rotatably meshed between the sun gear 41 and the ring gear 43, and a double planetary gear. The gear 44 is rotatably supported and the intermediate shaft 2 is connected via a multi-plate clutch 45.
6 and further connected to a second intermediate shaft 47 which is the output shaft of the planetary gear transmission 40.
, a hydraulic servo 48 that operates the multi-disc brake 42, and a hydraulic servo 49 that operates the multi-disc clutch 45. This planetary southeast gearbox 4 for forward and reverse switching
0, when the multi-disc clutch 45 is engaged and the multi-disc brake 42 is released, an old market gear with a reduction ratio l is obtained; when the multi-disc clutch 45 is disengaged and the multi-disc brake 42 is engaged, When it is, it becomes a reverse gear with a reduction ratio of 1.02. This reduction ratio of 1.02 during reversing is smaller than the reduction ratio of a normal automobile transmission during reversing, but in this example, the reduction ratio (for example, 24) obtained in a V-belt continuously variable transmission is smaller than that of a normal automobile transmission. Since deceleration is performed in the deceleration gear mechanism 23, which will be described later, an appropriate deceleration ratio can be obtained as a whole.

The speed gear mechanism 23 is intended to compensate for the fact that the speed range achieved by the belt-type continuously variable transmission 30 is lower than that achieved by a normal vehicle transmission, and is designed to reduce the speed difference between the input and output shafts. The torque is increased by changing gears with a reduction ratio of 1.45.

The differential gear 22 is connected to an axle (not shown) and provides final deceleration of 3.727:1.

FIG. 2 shows a hydraulic control circuit for controlling the continuously variable transmission 43 in the transmission device shown in FIG.

The hydraulic control circuit includes a hydraulic source 50, a hydraulic adjustment device 60, 'I
The engagement timing of the multi-disc brake 42 and the multi-disc clutch 45 in the Ji star gear shift Ja40 is controlled, and the N-
Shift control system to reduce shock during D and N-R shifts,
It consists of a transverse 70 and a torque ratio of 1,1 and 11 devices 80.

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 side. A torque ratio valve 66 that operates in conjunction with the movable flange 321 of the pulley 32 and supplies line pressure to the detent valve 64 according to the amount of displacement of the movable flange 321 and exhausts pressure from the output oil pressure feedback oil passage 9 provided in the throttle valve 65, and from the oil pressure source 50. It is composed of a regulator valve 61 that regulates the supplied hydraulic pressure and supplies it as line pressure to various parts of the hydraulic pressure adjustment device Z60.

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

The manual valve 62 moves the spool 621 to each position P, R, N, D, L as shown in FIG. As shown in Table I, the oil passage 1 to which line pressure is supplied communicates with the output oil passages 3 to 5.

Table ■ In Table ■, ◯ indicates a state of communication with oil passage 1, and × indicates that oil passages 3 to 5 are in a discharged pressure state.

The regulator valve 61 inputs the spool 6'l1, the detent pressure and the throttle pressure to the spool 611.
and a regulator valve plunger 612 for controlling the second output capotro 14 as the spool 611 is displaced.
The line pressure is output from the output port 16 to the oil passage 1. Oil is supplied from the port 1-614 through the oil passage 12 to the fluid coupling, oil cooler, and parts requiring lubrication.

The detent valve 64 includes a spool 641 that moves in conjunction with the throttle opening θ of the control valve of the carburetor 102 as shown in FIG.
≦θ, 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). Note that the spool 641 may be linked to the depression of the accelerator pedal operated by the driver.

The throttle valve 65 includes a spool 651 which is connected in series with the spool 641 of the detent valve via a spring 645 and has a spring 652 on the other side, and the throttle opening θ transmitted via the spool 641 and the spring 645. The above spool 6 moves according to fluctuations in
51, the opening area of the boat 653 that communicates with the oil passage 1 is adjusted, and the throttle pressure output oil passage 8 that communicates with the manual port 1B provided in the regulator valve 61 outputs heslosotor pressure. The spools 651 are branched from the oil passage 8 and have orifices 654 and 65, respectively.
The output hydraulic pressure is fed back to a land 656 and a land 657 having a larger pressure receiving area than the land 656 through the output hydraulic pressure feedback oil passages 9 and 10 provided with the output hydraulic pressure.

The torque ratio valve 66 includes a spool 662 linked to the movable flange 322 of the output side boob IJ32 via a connecting rod, and the movement iL of the movable flange 322 is j!
3 ≦L ≦1 a (torque ratio T 711<t z
≧T≧1+), 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 the drain boat 665 to discharge pressure. The amount of movement of the movable flange 322 is 12≦L<+! 3
When (t3≧T>L2), the spool 662 is located in the middle as shown in FIG. be pressured. Movement ff1L is O≦L, ≦e2(ta
≧T>t, ), the spool 662 is positioned on the right side of the figure as shown in FIG. Line pressure is supplied to the

In addition, the spool 662 is connected to the output pulley 3 which is in a rotating state.
The spool 662 slides in conjunction with the movable flange 322 of No. 2, but as shown in FIG. , the movement of the movable flange is not hindered, and it is possible to prevent wear and the like of sliding parts having large relative speeds.

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 oil passage 1 that supplies heline pressure
2. A pressure limiting valve 73 installed between the cough orifice 72 and the oil chamber 713, and a solenoid valve 74 that adjusts the oil pressure of the oil chamber 713 under the control of an electric control circuit to be described later. When the solenoid valve 74 operates to open the drain boat 741 and evacuate the oil chamber 713, the spool 712 of the shift control valve 71 is moved to the left in the figure by the action of the spring 711, and the planetary gear shift [40 Hydraulic servo 4 that operates plate clutch 45
9 and the oil path 14 that communicates with the hydraulic servo 48 that operates the multi-disc brake 42 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 drained. to be released. When the solenoid valve 74 is not operating, the drain boat 7
41 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, a downshift solenoid 84,
and an upshift solenoid 85. The torque ratio control valve 81 includes a spool 812 having a spring 811 on its back, and oil chambers 81 at both ends to which line pressure is supplied from the oil passage 1 through orifices 82 and 83, respectively.
5 and 816, an input boat 817 that communicates with the oil passage 1 to which line pressure is supplied, and whose opening area increases or decreases according to the movement of the spool 812, and a V-belt type continuously variable transmission 3.
An oil chamber 81 provided with an output port 818 that communicates with the hydraulic servo 313 of the input side boo U 31 via the oil passage 2.
9, 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.

The downshift solenoid 84 and the amplifier shift solenoid 85 are attached to an oil chamber 815 and an oil chamber 816 of the wholesale valve 81, respectively, and both are operated by the output of an electric control circuit to be described later. Oil chamber 8
15 and the oil chamber 816 are evacuated.

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 84 and the upshift solenoid 85 is shown.

901 is a shift lever switch that detects whether the shift lever is shifted to PSR, N, D, or L; 902 is a rotational speed sensor that detects the rotational speed of the input pulley 31; 903 is a vehicle speed sensor; 904 is a 905 is a speed detection processing circuit that converts the output of rotational speed sensor 902 into voltage; 906 is a vehicle speed detection circuit that converts the output of vehicle speed sensor 903 into voltage; 907 is a throttle opening detection processing circuit that converts the output of the throttle sensor 904 into voltage, 908 to 911 are input interfaces for each sensor, 912 is a central processing unit (CPU), and 913 is a solenoid valve 74.84.85
914 is a random access memory (RAM) that temporarily stores input data and parameters necessary for control; 915 is a clock; 916 is a is the output interface, 917 is the solenoid output driver, and the output interface 91
Upshift solenoid 85, downshift solenoid 84 and shift control solenoid 7
i to the operating output of 4.

Human power interfaces 908 to 911 and CPU 912,
ROM913, M914 to R, output interface 9
16 through a data bus 918 and an address bus 919.

Next, the torque ratio valve 66, detent valve 64, throttle valve 65, manual valve 62, and regulator valve 6
The operation of the hydraulic pressure adjustment device fi60 of this embodiment, which is comprised of 1, 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 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. The solid line in FIG. 7 shows the minimum line pressure necessary for a change in the torque ratio T between the input and output shafts when the engine is operated in a state that provides the best fuel efficiency, 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 ratios that can be achieved by using both brakes, so as shown by the dotted line, 20
It is desirable to set the line pressure to the line pressure indicated by the broken line, which is about % larger, and it is desirable to set the line pressure characteristic to be higher than the line pressure indicated by the dashed line even at the 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, throttle opening θ, and torque ratio of both pulleys (torque ratio between input and output shafts).

D position As shown in Table 1, 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 discharged. At this time, in the shift control mechanism 70, if the shift control solenoid 74 is in the OFF state and line pressure is supplied to the oil chamber 713, the spool 712 is positioned to the right, so that the oil passage 3 and the oil passage 1 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, and the vehicle becomes ready to move forward.

(1) When the torque ratio T increases, ≦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 the drain boat 665 to discharge pressure.

As a result, regardless of the throttle opening θ, the oil passage 7
Detent pressure (equal to line pressure) is not generated in the throttle valve 65, and the boat 664 of the torque ratio valve 66 connected to the oil passage 9 is closed, and the spool 651
Since the land 657 receives feedback pressure in addition to the land 656, the throttle opening θ changes as shown in Fig. 8(c).
A throttle pressure having the characteristics shown in is outputted to the regulating valve 61 and the regulator pulp plunger 612 via the oil passage 8.

As a result, the line pressure output from the regulating valve 61 is adjusted to (
f) and as shown in (e) of Figure 1O.

(2) When the torque ratio T is tt<'r≦t.

As shown in FIG. 5(B), the torque ratio valve 6G 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 oil passage 9 is exhausted and no feedback pressure is applied to the land 657 of the spool 651.
It is expressed by the characteristic curve shown in (d) in the figure. The line pressure at this time has characteristics shown in region (R) of FIG. 9 and (G) of FIG. 1O.

(3) 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 is expressed as (d) in FIG. 8, similar to (2) above. However, since the boat 663 opens and the oil passages 1 and 6 communicate with each other, the throttle opening θ is within the range of 0≦θ≦θ5%, and the detent valve 64
While the spool 641 is on the left side in the figure as shown in FIG. However, the throttle opening θ is 01% and 0510
When it is 0%, the spool 641 as shown in Fig. 4(B)
moves, 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 has a characteristic that changes stepwise at θ-θ1%, as shown in area (w) of FIG. 9 and (b) of FIG. 10.

■、Position At the manual valve 62, the oil passage 5 communicates with the oil passage 1.

Oil passage 3 and oil passage 4 are the same as position D.

(1) When the torque ratio T is t, ≦T≦L2.

When the throttle opening degree θ is 050509%, 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. When 0≦lOO%, the pressure in the oil passage 7 is exhausted through the drain boat 665 of the torque ratio valve as shown in oil passage 6 and No. 4 (B), and no detent pressure is generated. The throttle pressure is the same as in the D position. Therefore, the line pressure has the characteristics shown in No. 11 [2 (ru)].

(2) When the torque ratio T is t, <T≦t3.

The difference from the above (1) is that in the torque ratio valve 66, the oil passage 9 communicates with the drain boat 666 to exhaust pressure, and the throttle pressure output from the throttle valve 65 to the adjustment valve 61 via the oil passage 8 increases. In particular, the line pressure is expressed by a characteristic curve as shown in FIG. 11(h).

(3) When torque ratio 'l<tgokuT≦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 detent pressure regardless of the throttle opening, and inputs the detent pressure and the same throttle pressure as in (2) above. The regulating valve 61 outputs the line pressure shown in FIG. 11 (N).

As shown in R position table 1, in the manual valve 62, the oil passage 4 and the oil passage 5 communicate with the oil passage 1, and the oil passage 3 is depressurized. At this time, in the shift control mechanism 70, if the shift control solenoid 74 is in the OFF state and line pressure is being supplied to the oil chamber 713, the spool 712 is positioned to the left, so that the oil passage 4 is connected to the oil passage 14. are connected, and the oil passage 4
The line pressure supplied to is supplied to the hydraulic servo 48 of the reverse multi-disc brake 42 through the oil passage 14, and the vehicle enters the reverse state. Furthermore, since the line pressure is guided to the oil passage 5, the line pressure has the same characteristics as at the I7 position. In the R position, the torque ratio T at V-belt type continuously variable transmission [30] is used as the maximum T ``'' ta. For this reason,
Although it is not necessary to perform a speed change (deceleration) within the acid-N gear transmission 40, according to this embodiment, even when the torque ratio T is changed in the R position, the line pressure can be controlled in the same way as in the L position. is possible.

In the P position and N position manual valve 62, oil passages 3.4 and 5 are both exhausted, and since oil passage 5 is exhausted, 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 t3<'r≦t4 is
The characteristic curve shown in the figure (as in January, the throttle opening θ is set low at 1% or less is because the throttle opening θ is set low such as when idling).
If the line pressure is set high in an operating situation where the oil pressure is small and the pump discharge is low, it will be difficult to maintain the line pressure when the oil temperature is high and oil leaks from various parts of the hydraulic circuit. This is because a decrease in the amount of oil supplied to the cooler causes the oil temperature to rise further, which is likely to cause trouble. In addition, when the manual valve 62 is shifted to the G and R shift positions, the torque ratio T is [1≦T≦L] as shown in the characteristic curves (H) and (L) in FIG.
The reason why the line pressure is set high in the operating condition where the throttle opening θ is 83% or less in the range 2 is that a relatively high oil pressure is required even at a low throttle opening during engine braking. The required oil pressure at that time is indicated by a line at a point ui in FIG. 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 and fuel consumption rate.

Next, the shift system controlled by the electric control circuit g90 explained in FIG. In this embodiment, the operation of the III mechanism 70 and the torque ratio control device 80 will be explained with reference to the program flowcharts shown in FIGS. An example of controlling the rotation speed N is shown.

Generally, when operating an engine at its best fuel efficiency,
The vehicle is operated according to the best fuel efficiency power line shown in the broken line in the graph of FIG. In this Figure 12, the horizontal axis is the engine rotation speed (rp
m), the vertical axis shows the torque (kg-m) of the engine output shaft, and the best fuel efficiency power line is obtained as follows. That is, the constant fuel consumption rate curve of the engine (shown by the solid line in FIG. 12)
The unit is g/ps-h) and the equal horsepower curve shown by the two-dot chain line (
The unit is ps), then the fuel consumption rate Q at point A in the diagram
Cg/pS ' h), and horsepower is P (ps),
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 each throttle opening θ shown in FIG. , find the best fuel efficiency fluid coupling output line on the fluid coupling output performance curve as shown in FIG. 16 from the fluid coupling performance curve shown in FIG. 14 and the engine equivalent fuel efficiency rate curve shown in FIG. 15. Can be done. The 17th and 1 are the 16th
The best fuel efficiency fluid coupling output line shown in the table has been replaced with the relationship between throttle opening and fluid coupling output rotation speed. This fluid coupling output rotation speed directly becomes the input pulley rotation speed in the continuously variable transmission of this embodiment.

In the continuously variable transmission of this embodiment, the input side boolean 1J31 and the output side pulley 32 are Controls the gear ratio.

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 bag W80, so that the actual input pulley rotation speed is brought to the optimum input pulley rotation speed. FIG. 18 shows an overall flowchart of input side pulley rotation speed control.

After reading the throttle opening θ by the throttle sensor 904 (<921), the shift lever switch 90
1, the shift lever position is determined (922). As a result of the determination, if the shift lever is in the P position or N position, both solenoid valves 84 and 85 are turned OFF by a P position or N position processing subroutine (not shown in FIG. 19).
The L (931), P or N state is stored in the RAM 914 (932). As a result, the large force pulley 31 is brought into a neutral state. When shifting from P position or N position to R position, and from N position to D
When the position changes, shift shock control processing is performed to alleviate the shift shock associated with the N-R shift and N-D shift, respectively (940, 95
0). Shift shock control is performed by applying pulses with gradually decreasing pulse widths to the shift control solenoid valve 74 shown in FIG. 20 (hereinafter referred to as duty control). By controlling the duty, a hydraulic pressure P adjusted according to the duty is generated in the oil chamber 713 of the shift control valve 71.

The shift control '4'n 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 engagement pressure of the clutch and brake.

When mitigating engagement shock during N-D shift and N-R shift, hydraulic servo 48 or hydraulic servo 4
9, the rise of the oil pressure P or PC is rolled as shown in the oil pressure characteristic curve shown in FIG.
complete the bond. FIG. 25 shows 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. . Duty (%) is given by the following formula.

Duty (%) = Solenoid operating time The solenoid pressure shown in Fig. 25 is the shift control 2Bl valve 71.
The hydraulic pressure P or PC supplied to the hydraulic servo 48 or 49 shown in FIG. 26 is obtained.

In this embodiment, as shown in FIG. 24(A), the pressure receiving areas of the lands provided on the spool 712 of the shift control valve 71 are S1, So, and S in order from the left side in the figure.

, Sl, the elastic force of the spring 711 is F, and the oil chamber 713
Assuming that P is the hydraulic pressure of the shift control valve, the hydraulic pressures PC and Pb supplied to the hydraulic servo 49 of the multi-disc clutch 45 that is engaged during forward movement and the hydraulic servo 48 of the multi-disc brake 42 that is engaged during reverse movement are respectively From the 0th equation and the 0th equation, which are the hydraulic equilibrium equations of 71, it is given as follows.

When moving forward P5 xS, = pcXSz +Fs+ ■
When reversing S z S P5 xS, ""Pb X (s+ St)
10F, 1 ■ S + S z S IS t Also, the valve body 731 inserted in the pressure limiting valve 73
Suppose that the pressure receiving area is S, and the elastic force of the spring 732 placed behind the 3.8M valve body 731 is F, 2, the pressure limiting valve 73 operates at the maximum pressure pffimit of P3 according to the hydraulic balance equation (2). .

p j! i+n1LX 33-Fsz
■pl1iLIit=Fsx/Sx At this time, Pe and P are limited to the maximum pressure pcJimit, pb 6in+it according to the 0th equation and the 0th equation.

Sl when moving forward L when moving backward pb 1 imit = -p 12 in+it
S-I SI S + S z Returning to FIG. 18, following the N-D shift shock control processing 950, the input-side pulley rotation speed sensor 902 reads the actual input-side pulley rotation speed N ( 923), then it is determined whether the throttle opening degree θ is O or not (924), and when θ≠0, the data is stored in the ROM 913 in advance according to the subroutine shown in FIG.
Set the best fuel economy input side pulley rotation speed N9 corresponding to the throttle opening θ shown in FIG. 17 (960
) Therefore, set the storage address for the input pulley rotation speed N'' data corresponding to the throttle opening degree (961), and read the No data from the set address (962).
), transfer the read data of N1 to R for data storage M91
4 (963).

Next, the actual input side pulley rotation speed N and the best fuel economy input side pulley rotation speed N0 are compared (927). When N=N'', an OFF command is issued for both solenoid valves 84 and 85 (920).

When θ=O and the throttle is fully closed, it is determined whether the shift lever is set to the D position or the L position in order to determine the necessity of engine braking (
926), and engine braking processing 970 or 980 is performed as necessary. D position engine brake processing 97
0, as shown in FIG. 22, the vehicle speed V is read by the vehicle speed sensor 903 (971), the acceleration α at that point is calculated (972), and then the acceleration α is an appropriate acceleration A for the vehicle speed. It is determined whether or not (973). α〉
In the case of A, in order to perform downshift control 974, return after setting No to a value larger than N, and α≦
In the case of A, the best fuel efficiency input pulley rotation speed N'' corresponding to the throttle opening degree O is set in N1 (975), and then the process returns.

The relationship between vehicle speed and appropriate acceleration A is determined by experiment or calculation for each vehicle, and is shown in (Fig. 22).
It is shown in the graph of B).

In the engine brake processing 980 for the L position, as shown in FIG. 23, after reading the vehicle speed V (981), an operation is performed to calculate the torque ratio T from the vehicle speed V and the input pulley rotation speed N using the following formula. (982).

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. Next, the current torque ratio T is the vehicle speed■
It is determined whether the torque ratio T' is greater than the torque ratio T' at which safe and appropriate engine braking can be obtained (983).
, TNT” so that a downshift is performed.
is set to a value larger than N (984), and 'F≧T0
In this case, set N1 to a value equal to N (985).
Return. The torque ratio T0 that provides safe and appropriate engine braking for each vehicle speed is determined by experiment or calculation for each vehicle, and is shown in Figure 23 (
It is shown in the graph of B).

Next, the operation of the torque ratio jt+II control device 80 will be explained with reference to FIG. 27.

When the vehicle is traveling at a constant speed, the solenoid valves 84 and 85 controlled by the output of the electric control circuit are turned off, as shown in FIG. 27(A). 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. Spool 812
is moved to the left in the figure because of the pressing force P exerted by the spring 811. When the spool 812 is moved to the left and the oil chamber 815 and the drain port 813 communicate with each other, the pressure P2 is exhausted, so the spool 812 is moved to the left to connect the oil chamber 815 and the drain port 813.
is moved to the right in the diagram. When spool 812 is moved to the right, drain port 813 is closed. Therefore, in this case, as shown in FIG. 27, the spool 812 is provided with a flat notch 812b at the land edge between the drain port 813 and the spool 812, so that the spool 812 can be moved in a more stable state as shown in FIG.
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 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 pulley 31 is gradually expanded and the torque ratio T changes in the direction of increasing. Therefore, as shown in FIG. 27(A), when the spool 812 is in equilibrium, the drain port 814 is closed and
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. Also the 29th
It is clear that the same function can be achieved even if the oil passage 822 having an orifice 821 connects the oil passage 1 and the oil passage 2 instead of the notch 812a as shown in the figure.

At the time of amplifier shift, the solenoid valve 85 is turned on by the output of the electric control circuit as shown in FIG. 27(B). 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 port 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 bow) 818, so the oil pressure of hydraulic servo 313 increases, the input pulley 31 operates in the direction of closing, and the torque ratio T decreases. Decrease. Therefore, by controlling the ON time of the solenoid valve 85 as necessary, the torque ratio is reduced by a desired amount to perform an upshift.

During a downshift, the solenoid valve 84 is turned on by the output of the electric control circuit as shown in FIG. 27(C), and the pressure in 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 port 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 hydraulic servo 312 of the input side (drive side) pulley 31 is supplied with the output hydraulic pressure of the torque ratio control valve 81, and the hydraulic servo 323 of the output side (driven side) boolean U32 is supplied with line pressure. Then, the oil pressure of the input side hydraulic servo 312 is set to P.

, If the oil pressure of the output side hydraulic servo 322 is Po, then P,
/Pi has characteristics as shown in the graph of FIG. 28 with respect to torque ratio T, for example, throttle opening θ=50%,
If you release the accelerator from the state where you are driving with torque ratio T = 1.5 (point a in the figure) and set θ = 30%, P0/r
', is maintained as it is, the torque ratio T = 0.87
If you move to the point in the figure and conversely maintain the state of torque ratio T = 1.5, will the input side boob be suppressed? The value of P, /Pi is increased by the output of the torque ratio control mechanism 80 which increases by 2U.
Change to point value. By controlling the values of P, /I)i 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 according to the present invention, and Fig. 3 is a diagram for explaining the operation of a manual valve. , Figure 4 is a diagram for explaining the operation of the detent valve and throttle valve, Figure 5 is a leap diagram for explaining the operation of the torque ratio valve, and Figure 6 is an embodiment of the electric control circuit according to 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 control device of the present invention. Figure 12 is a diagram showing the best fuel consumption power line of the engine, Figure 13 is a diagram showing the output performance characteristics of the engine, and Figure 14 is a diagram showing the performance of the fluid transmission machine (?4). Figure 15 is a diagram showing the equal fuel consumption rate curve of the engine, Figure 16 is a diagram showing the best fuel efficiency fluid coupling output curve, and Figure 17 is a diagram showing the characteristics of the best fuel efficiency fluid coupling output rotation speed. Figures 18, 19, 21, 22 (A), and 23 (A) are diagrams for explaining the flow of processing in the electric control circuit.
Figure 0 is a diagram for explaining the action of the solenoid valve, No. 22.
Figure (B) shows the set acceleration, Figure 23 (B) shows the set torque ratio, and Figure 24 (A) shows the shift control a t
14, FIG. 24(B) is a diagram showing the characteristics of the hydraulic pressure supplied to the input and output side hydraulic servos, FIG. 25 is a diagram showing the characteristics of the solenoid pressure, and FIG. A diagram showing the characteristics of the output oil pressure of the shift control valve, FIG. 27 is a diagram to explain the operation of the torque ratio control device, and FIG. 28 is a diagram showing the relationship between the torque ratio and the pressure ratio of the input/output side hydraulic servo. , FIG. 29 is a block diagram showing another embodiment of the torque ratio control device, and FIG. 30 is a schematic diagram of a conventional V-belt type continuously variable transmission for a vehicle. 30...Continuously variable transmission, 214...Human power shaft, 26...
・Output shaft, 31...Input side pulley, 32...Output side pulley, 313.323...Hydraulic servo, 33...
Drive band, 90... Electric control circuit, 902... Input side pulley rotation speed, 903... Vehicle speed sensor, 904...
... Throttle sensor, 84... Downshift solenoid valve, 85... Upshift solenoid valve, 81.
...Torque ratio control valve. Fig. 1 Fig. 3 Fig. 4 (A) (Snake) Fig. 5 (B) Fig. 6 Fig. 7 (K9/Cm2) Throttle opening θ Fig. 9 Large η I'l Houli l; iL (mm) Figure 12 Figure 13 Figure 14 Figure 15 Figure 16 Figure 17 Figure 18 Figure 19 Figure 21 Figure 20 Figure 22 (A) (Snake) Dou Figure 23 (A) CB) 4L Figure 24 (A) (已) Pcir=+tPb (kg/crn”)PS
River-r%8(gu〈 Figure 28 Figure 29 Figure 30

Claims (7)

[Claims] (1) Consisting of an input-side pulley and an output-side pulley that are attached to the input and output shafts and have variable effective diameters, and a drive band stretched between these pulleys, the effective diameter of the pulleys can be adjusted. In a vehicle V-belt continuously variable transmission that controls the torque ratio between input and output shafts, there is a means for detecting vehicle operating conditions such as throttle opening, vehicle speed, input pulley rotation speed, etc., and a signal from the detecting means. means for setting a corresponding shift target value, and comparing these shift target values and the detected value,
an electric control circuit that outputs a signal based on the comparison result;
and a torque ratio control means for adjusting the torque ratio according to the output signal of the electric control circuit, and the means for setting the gear shift target value includes a plurality of means for setting gear shift target values having mutually different characteristics, and the means for setting the gear shift target value has different characteristics. 1. A speed change control device for a V-belt type continuously variable transmission for a vehicle, characterized in that a speed change target value is set by selecting one of the plurality of means in accordance with the above. (2) A patent characterized in that the means for detecting the operating state of the vehicle detects an engine brake state, and at least one of the plurality of means for setting the gear shift target value is selected in the engine brake state. A speed change control device for a V-belt type continuously variable transmission for a vehicle according to claim 1. (3) The means for determining the state of the engine brake determines a state where the throttle opening is zero or a state where the throttle opening is zero and the shift range is L. Shift control device for V-belt continuously variable transmission for vehicles. (4) A shift control device for a V-belt type continuously variable transmission for a vehicle according to claim 2 or 3, wherein the shift target value for engine braking is an acceleration calculated based on the vehicle speed. . (5) The speed change control device for a V-belt type continuously variable transmission for a vehicle according to claim 4, wherein the speed change target value for engine braking is set in the D range. (6) The V-belt type vehicle according to claim 2 or 3, wherein the target speed change value for engine braking is a torque ratio calculated from the vehicle speed and the rotational speed of the input pulley. Speed change control device for gear transmission. (7) The speed change control device for a V-belt type continuously variable transmission for a vehicle according to claim 6, wherein the speed change target value for engine braking is set in the L range.