JP3514052B2 - Continuously variable transmission - Google Patents

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a continuously variable transmission that produces torque circulation in combination with a planetary gear, and more particularly to a continuously variable transmission that is suitable for mounting on an automobile. The present invention relates to an continuously variable transmission (IVT).

[0002]

2. Description of the Related Art Recently, an automatic transmission incorporating a belt type continuously variable transmission (CVT) has been attracting attention as a transmission of an automobile due to demands such as improvement of fuel consumption rate and improvement of driving performance.

Conventionally, as disclosed in Japanese Patent Application Laid-Open No. 6-331000, a continuously variable transmission mechanism, a constant speed mechanism, and a planetary gear mechanism are provided, and the planetary gear mechanism is used to drive the continuously variable transmission mechanism and the constant speed mechanism. A continuously variable transmission has been devised in which power is combined to generate power (torque) circulation in a continuously variable transmission mechanism to amplify a gear shift range. The continuously variable transmission transmits the engine output to the carrier via the constant speed mechanism and to the sun gear via the continuously variable transmission mechanism and the first (low) clutch or the one-way clutch. As the speed circulation mechanism generates torque circulation and its gear ratio changes from small (O / D) to large (U / D), the gear ratio of the continuously variable transmission output shaft from the ring gear goes backward → ∞ (output rotation speed 0) → large forward (U / D) → small forward (O / D), further releasing the first clutch or the one-way clutch and engaging the second (high) clutch to provide a continuously variable transmission mechanism. Is directly transmitted to the output shaft, and as the gear ratio of the continuously variable transmission mechanism becomes large (U / D) to small (O / D), the rotation ratio of the output shaft also becomes large (U / D).
The speed is changed from D) to small (O / D).

The continuously variable transmission that circulates the torque described above geometrically rotates the output shaft by setting the speed ratio of the continuously variable transmission mechanism to a predetermined value determined by the gear ratio of the planetary gear mechanism. There is a gear neutral position where the number is 0, and thus theoretically holds even without a starting device such as a starting clutch.

[0005]

In the continuously variable transmission with torque circulation described above in the gear neutral position,
Due to the infinite spread of the torque ratio, the torque ratio becomes extremely large even in the vicinity thereof. Therefore, it is necessary to limit the input torque to the continuously variable transmission by being restricted by the allowable transmission torque capacity of the transmission element of the continuously variable transmission, particularly the belt of the CVT.

Generally, a continuously variable transmission mounted on a vehicle is
An internal combustion engine is used as a power source, but in the vicinity of the gear neutral position where the torque amplification ratio is maximized (when the vehicle starts), the throttle opening is extremely low (for example, about 5%) due to the limitation of the input torque. Can be suppressed to.
Even in the case of engine output by the extremely low opening throttle, the continuously variable transmission can handle the required torque because the torque amplification ratio is maximized. The characteristic equal throttle line is declining to the right, and the engine speed is also very low. Therefore, even if the driver steps on the accelerator pedal to accelerate, the torque is sufficiently output, but the engine speed remains low, and the CVT is upshifted (shifted in the U / D direction). If the engine speed is increased, torque will not be output sufficiently.

This is a starting device which has been generally used in the past.
For example, the characteristics of the torque converter, that is, "When the accelerator pedal is depressed, the engine speed increases accordingly.
In addition, the input torque of the transmission becomes large, which makes the driver feel uncomfortable.

Therefore, the present invention suppresses the throttle of the engine to a very low opening at the time of starting and limits the input torque to the continuously variable transmission in accordance with the torque amplification ratio by the torque circulation. Is appropriately controlled to solve the above problems.

[0009]

[0010]

[Summary of the present invention according to claim 1, an input shaft interlocked with the output shaft of the engine (2) (3)
And an output member (21) interlocking with the wheel (5), a first rotary member (7) interlocking with the input shaft, a second rotary member (9), and a rotation ratio of these rotary members. It has a continuously variable transmission (11) having shift operating means (7c, 9c) (7c, 9d) (7e, 9e) and at least first, second and third rotating elements (19c, 19s, 19r). Then
Based on the change of the rotation ratio (I _P ) of the continuously variable transmission (11), the torque transmission direction is changed between the rotary members and the output torque direction of the output member (21) is changed. The first rotating element (19c) to the input shaft (3), the second rotating element (19s) to the second rotating member (9), and the third rotating element (19r).
In a continuously variable transmission (1) comprising a planetary gear (19) interlocking with the output member (21), when starting from a neutral position where the rotation of the output member is 0, An accelerator operating means for manually operating the engine within a throttle opening within which the output torque (Te) of the engine is within a limit torque set according to the rotation ratio of the continuously variable transmission. Of the electronic throttle control means (Q) for controlling the throttle opening (θ ^* ) so that the target rotation speed corresponds to the manipulated variable (θd) of the engine, and the output torque of the engine by the electronic throttle control means. The continuously variable transmission (1
1) is provided with engine load control means (R) for controlling the shift operation means so that an urging force (ΔP) that deviates from the neutral position (GN) by a predetermined amount acts. In the continuously variable transmission (Fig. 21,
(See FIG. 22).

According to a second aspect of the present invention, the engine load control means (R) controls the shift operation means so that the engine speed becomes a target speed (N _ei ^* ). The continuously variable transmission according to claim 1 (see FIGS. 21 and 2).
2).

According to a third aspect of the present invention, the electronic throttle control means (Q) controls the limit torque based on the throttle opening by the electronic throttle control means when the accelerator operation means is operated to the maximum. 3. The continuously variable transmission according to claim 1, wherein the throttle opening is controlled so that the output torque (Te) is set in accordance with the ratio of the operation amount of the accelerator operating means (Fig. 24,
(See FIG. 25).

According to a fourth aspect of the present invention, in the continuously variable transmission (11), first and second pulleys (7, 9), a belt (10) wound around these pulleys, and A hydraulic actuator (7) that applies an axial force to both pulleys.
c, 9c) (7c, 9d) (7e, 9e), the engine load control means (R) operates the hydraulic actuator to operate the belt type continuously variable transmission. The continuously variable transmission according to claim 2 , wherein (11) is controlled such that the engine speed becomes a target speed (N _ei ^* ).

[Operation] Based on the above configuration, the input shaft (3)
The engine output from the continuously variable transmission (1
1) is appropriately speed-changed and transmitted to the second rotary element (19s) of the planetary gear (19), and the constant speed rotation is transmitted to the first rotary element (19c). It is combined in (19) and transmitted from the third rotating element (19r) to the drive wheel via the output member (21). At this time, torque circulation is generated so as to sandwich the neutral position (GN) where the rotation of the output member becomes 0, and the rotation direction of the output member is switched between normal rotation and reverse rotation by the rotation ratio of the continuously variable transmission (11). .

The electronic throttle control means (Q) keeps the rotational ratio (I _P ) near the neutral position (GN) within the limit torque set according to the rotational ratio of the continuously variable transmission when the vehicle starts. The throttle opening is controlled by. Under normal running conditions, the engine speed changes significantly due to a sudden change in running resistance.Therefore, the throttle opening of the engine should be above the equal throttle opening of the engine output characteristics so that the limit torque is not exceeded even if it changes. Although it is necessary to limit the maximum torque as a reference, the throttle opening does not have to be limited based on the maximum torque because the vehicle is not affected by external influences such as the running resistance when the vehicle starts.
Therefore, the throttle opening of the engine is controlled by the electronic throttle control means (Q) so that the target rotation speed is within the above-mentioned limit torque according to the operation amount of the accelerator operation means.

At the same time, the engine load control means (R) is
An urging force (Δ) that causes the continuously variable transmission (11) to deviate from the neutral position (GN) by a predetermined amount according to the engine output torque by the electronic throttle control means (Q).
P) is operated to control the gear shift operation means, for example, the load acting on the engine is controlled so that the engine speed becomes the target speed (N _ei ^* ).

The reference numerals in parentheses are for comparison with the drawings, but do not limit the structure of the present invention.

[0018]

[0019]

According to the present invention of claim 1 , at the time of starting
With the ability to set the engine speed of
In addition, by applying an urging force to the continuously variable transmission so that the continuously variable transmission deviates from the neutral position by a predetermined amount according to the output torque of the engine to control the engine load, the creep at vehicle start The torque can be set arbitrarily. Therefore, it is possible to obtain stall characteristics (rotational speed, torque) suitable for a starting device such as a conventional torque converter, electromagnetic clutch, manual clutch, etc., and it is possible to smoothly start.

According to the second aspect of the present invention, by changing the engine load by controlling the urging force of the continuously variable transmission, the engine speed is stabilized at the target value speed and the engine speed is accurate. The desired stall characteristics consisting of

According to the third aspect of the present invention, the engine output torque is controlled in accordance with the ratio of the operation amount of the accelerator operating means to the maximum limit torque when the accelerator operating means is operated to the maximum. It is possible to obtain the engine output torque that matches the operation feeling of.

According to the fourth aspect of the present invention, the engine speed can be stabilized at the target speed by the simple and highly reliable control using the belt type continuously variable transmission and the hydraulic actuator.

[0023]

BEST MODE FOR CARRYING OUT THE INVENTION Embodiments of the present invention will be described below with reference to the drawings. The in-vehicle automatic continuously variable transmission 1 includes an engine crankshaft 2 as shown in FIG.
The first axle 3, the second axle 4, the third axle aligned with the front axle
Fourth shaft 6 including shaft 5 (a, b) and counter shaft
And a primary (first) pulley 7 is arranged on the first shaft 3, and a secondary (second) pulley 9 is arranged on the second shaft 4. 7,
A belt 10 is wound around 9 to form a belt type continuously variable transmission 11.

Further, the first shaft 3 is directly connected to the engine crankshaft 2 via a damper device 12 that absorbs torque fluctuations of the engine to form an input shaft.
Is composed of a fixed sheave 7a of the primary pulley 7 and a shaft 3a spline-fitted to the boss portion 7a ₁ of the fixed sheave 7a. Then, the constant on the input shaft 3 and the output-side member 15 with the shaft 3a input side member 13 of the low clutch C _L is fixed to configure are rotatably supports, and output-side member 15 The primary side sprocket 18 that constitutes the high speed transmission 16 is integrally connected. An oil pump 17 is connected to a fixed sheave 7a of a primary pulley 7 constituting the input shaft 3, and a movable sheave 7b is supported on the fixed sheave 7a by a hydraulic actuator 7c described later so as to be movable in the axial direction. There is.

The second shaft 4 is composed of a fixed sheave 9a of the secondary pulley 9, and a movable sheave 9b is supported on the fixed sheave 9a by a hydraulic actuator 9c so as to be movable in the axial direction. Further, a high clutch C _H and a planetary gear 19 are arranged on the second shaft 4, and a secondary side sprocket 20 and an output gear (output member) 21 are rotatably supported.

The planetary gear 19 is a sun gear 19
s, a ring gear 19r, and a carrier 19 that rotatably supports a pinion 19p that meshes with these gears.
It consists of a single pinion planetary gear with c.
The sun gear 19s is connected to the fixed sheave 9a of the secondary pulley 9 forming the second shaft 4 to form a second rotating element, and the ring gear 19r is connected to the output gear 21.
To form a third rotating element, said carrier 1
9c is connected to the secondary side sprocket 20
Constitutes the rotating element of. Further, a winding body 22 such as a chain such as a silent chain or a roller chain or a timing belt is wound around the primary side and secondary side sprockets 18 and 20. Further, a high clutch C _{H is provided} between the sun gear 19s and the ring gear 19r.
Is intervening.

The output gear (output member) 21 meshes with a large gear 23a of the counter shaft 6 constituting the fourth shaft, and the small gear 23b of the shaft meshes with a ring gear 24 of the differential device 25. The differential device 25 outputs differential rotations to the left and right axle shafts 5a and 5b forming the third shaft.

The hydraulic actuators 7c and 9c of the primary pulley 7 and the secondary pulley 9 have partition members 45 and 46 and cylinder members 47 and 49 fixed to the fixed sheave boss portions 7a ₁ and 9a ₁ , respectively, and a movable sheave 7b. , 9b The drum member 5 fixed to the back surface
0,51 and the second piston members 52,53, and the partition members 45,46 are the second piston members 52,5.
3 and the second piston members 52 and 53 are fitted into the cylinder members 47 and 49 and the partition member 4 in an oil-tight manner.
5, 46 are oil-tightly fitted to the first hydraulic chamber 5 respectively.
5, 56 and the second hydraulic chambers 57, 59 have a double piston (double chamber) structure.

Then, the hydraulic actuators 7c, 9
In the first hydraulic chambers 55 and 56 in c, the back surfaces of the movable sheaves 7b and 9b form piston surfaces, and the effective pressure receiving areas of the piston surfaces are equal on the primary side and the secondary side. Further, the primary side and secondary side fixed sheave bosses 7a ₁ and 9a ₁ respectively have oil passages 32 and 33 communicating with the first hydraulic chambers 55 and 56 and oil passages 35 communicating with the second hydraulic chambers 57 and 59, respectively. 36 is formed, and a preload spring 65 for urging the movable sheaves 7b, 9b on the primary side and the secondary side toward the fixed sheaves 7a, 9a, respectively.
66 is contracted.

Next, the operation of the continuously variable transmission 1 will be described with reference to FIGS. 1, 2 and 3. The rotation of the engine crankshaft 2 is transmitted to the input shaft 3 via the damper device 12. In the D mode, in the low mode in which the low clutch C _L is connected and the high clutch C _H is disconnected, the rotation of the input shaft 3 is transmitted to the primary pulley 7 and the primary side sprocket 18 Is transmitted to the carrier 19c of the planetary gear 19 via the constant speed transmission device 16 composed of the winding body 22 and the secondary side sprocket 20. On the other hand, the rotation of the primary pulley 7 is controlled by the hydraulic actuator 7 described later.
By appropriately adjusting the pulley ratios of the primary and secondary pulleys by c and 9c, the speed is continuously changed and transmitted to the secondary pulley 9, and the rotation of the pulley 9 is transmitted to the sun gear 19s of the planetary gear 19.

In the planetary gear 19, as shown in the velocity diagram of FIG. 2, the carrier 19c to which the constant speed rotation is transmitted via the constant speed transmission device 16 serves as a reaction force element, and the belt type continuously variable transmission ( The continuously variable rotation from the CVT 11 is transmitted to the sun gear 19s, and the rotations of the carrier and the sun gear are combined and transmitted to the output gear 21 via the ring gear 19r. At this time, since the output gear 21 is connected to the ring gear 19r, which is a rotating element other than the reaction force supporting element, the planetary gear 19 causes torque circulation, and the sun gear 19s and the carrier 19c rotate in the same direction. , The output shaft 5 rotates forward (L
o) and rotation in the reverse (Rev) direction. That is, based on the torque circulation, when the output shaft 5 is rotated in the forward (forward) direction, the belt-type continuously variable transmission 11 transmits torque from the secondary pulley 9 to the primary pulley 7 to reverse the output shaft (reverse). In the direction rotation state, the primary pulley 7
Torque is transmitted from the secondary pulley 9 to the secondary pulley 9.

In the high mode in which the low clutch C _L is disengaged and the high clutch C _H is engaged, the transmission to the planetary gear 19 via the constant speed transmission 16 is cut off, and the planetary gear 19 is disconnected. Is brought into an integrated rotation state by engagement of the high clutch C _H. Therefore, the rotation of the input shaft 3 is limited to the belt type continuously variable transmission (CVT) 1
1 and the high clutch C _H to the output gear 21. That is, the CVT 11 transmits power from the primary pulley 7 to the secondary pulley 9. Further, the rotation of the output gear 21 is caused by the rotation of the gears 23 a, 2 of the counter shaft 6.
Is transmitted to the differential device 25 via 3b,
It is transmitted to the left and right front wheels via the left and right axle shafts 5a and 5b.

As shown in the velocity diagram of FIG. 2, the output torque diagram of FIG. 4, and the output rotational speed diagram of FIG. 5, the belt type continuously variable transmission (hereinafter referred to as CVT) 11 operates in the low mode. When it is at the limit (O / D end) in the speed increasing direction (position a in FIG. 2), the ring gear 1 is rotated with respect to the rotation of the carrier 19c at a constant rotation speed based on the maximum rotation of the sun gear 19s.
9r is reversed, and reverse rotation (REV) is transmitted to the output gear 21. When the CVT 11 shifts in the deceleration (U / D) direction, the rotational speed of the reverse rotation decreases, and the rotational speed of the output gear 21 at the predetermined pulley ratio determined by the gear ratio of the planetary gear 19 and the constant speed transmission 16. Becomes the neutral position (NEU) where becomes zero. Further, when the CVT 11 shifts in the deceleration direction, the ring gear 19r is switched to the forward rotation direction, and the forward rotation, that is, the forward rotation is transmitted to the output gear 21. At this time, as is clear from FIG. 4, in the vicinity of the neutral position NEU,
The torque of the output gear 21 diverges infinitely.

Then, the CVT 11 is in the deceleration direction (U / D).
At the end, the high clutch C _H is connected and switched to the high mode. In the high mode, the output rotation of the CVT 11 is transmitted to the output gear 21 as it is, so the speed diagram of FIG. 2 becomes parallel lines as indicated by b. Then, this time, as the CVT 11 is shifted in the speed increasing (O / D) direction, the rotation of the output gear 21 is also changed to the speed increasing direction, and the transmission torque is reduced accordingly. In addition, λ in FIG. 2 is the number of teeth Zs of the sun gear and the number of teeth Z of the ring gear.
It is the ratio to r (Zs / Zr).

In the parking range P and the neutral range N shown in FIG. 3, both the low clutch C _L and the high clutch C _H are disengaged, and the power from the engine is disengaged. At this time, in the parking range P, the differential device 25 is locked and the axle 5
a and 5b are locked.

Next, the hydraulic control mechanism according to this embodiment will be described with reference to FIG. The hydraulic control mechanism 70
Is a primary regulator valve 71, a ratio control valve 72, a downshift relief valve 73,
Manual valve 75, low-high control valve 7
6, a low clutch relief valve 77, and a clutch modulation valve 79, and further includes a ratio sensing valve 80 as a restricting means, a sensor shoe 81 as a pulley ratio detecting means, and an interlock rod 82 as a locking means. I have it.

The sensor shoe 81 is the primary pulley 7
It is slidably supported by a guide member 83 arranged in parallel with the axis of. 2 from the sensor shoe 81
Book connecting portions 81b and 81c are projected, one connecting portion 81b is engaged with the movable sheave 7b of the primary pulley 7, and the other connecting portion 81c is engaged with the ratio sensing valve 80 described above. ing. Therefore, when the movable sheave 7b moves in the O / D direction or the U / D direction along the axis, the movement amount is transmitted to the ratio sensing valve 80 as it is via the sensor shoe 81.

Further, the sensor shoe 81 has a recess 81a.
Is formed, and the base end 82a of the interlock rod 82 is disengaged from the recess 81a. The interlock rod 82 is arranged so as to pass through the valve body, and a tip end portion 82b thereof is engaged with and disengaged from the recesses 76a and 76b of the low-high control valve 76.
In addition, in FIG. 6, the interlock rod 82 is
Although the base end portion 82a and the front end portion 82b are shown separately, in reality, they are integrally formed. Further, when the base end portion 82 a of the interlock rod 82 is engaged with the recess portion 81 a of the sensor shoe 81, the tip end portion 82 b has the recess portion 7 of the low-high control valve 76.
6a, 76b is abutted on the surface of the low / high control valve 76 without being engaged with the base end portion 8
When 2a comes off the concave portion 81a of the sensor shoe 81 and is in contact with the surface of the sensor shoe 81, the tip portion 82b
Is the recesses 76a, 7 of the low-high control valve 76.
It is configured to engage with any of 6b.

In the neutral state, the above-mentioned hydraulic control mechanism 70 supplies hydraulic pressure to both the primary and secondary side first hydraulic chambers 55 and 56, and at the same time, to both the second hydraulic chambers 5 and 56.
When the hydraulic pressures of 9 and 57 are released and the vehicle starts in the forward direction from the neutral state, the hydraulic pressure is supplied to the second hydraulic chamber 59 on the secondary side to shift the CVT 11 in the U / D direction, and to reverse. At the time of starting the direction, the second hydraulic chamber 57 on the primary side is supplied to shift the CVT in the O / D direction. At this time, in order to reliably prevent the low-high control valve 76 from being switched due to a malfunction of the computer and starting in the reverse direction in the D range, for example, in the predetermined forward range of the CVT, the ratio sensing valve 80 downshifts. In addition to prohibiting, switching of the low-high control valve 76 is mechanically restricted by the interlock rod 82. Specifically, with a predetermined pulley ratio B (for example, 1.3) slightly larger than 1.0 in the forward range shown in FIG. 5 as a boundary, a region (U / D side) where the pulley ratio is larger and a region where the pulley ratio is smaller than this ( O / D side) and control are changed respectively. By changing the control, the D range low mode and the R range are set in the range of the predetermined pulley ratio B or less.
The downshift in the range is prohibited, and the jump from the D range H mode to the D range L mode and the R range is prohibited.

Next, the operation of the hydraulic control mechanism 70 having the above configuration will be described with reference to FIG. In the following,
(1) D range low mode, (2) D range high mode, (3) R (reverse) range, (4) N (neutral), and P (parking) range will be described in this order. First,
In any of the modes (1) to (4), as shown in FIG. 6, the hydraulic pressure from the oil pump 17 is appropriately regulated by the primary regulator valve 71 and output from the output port v, and the primary And the first hydraulic chambers 55, 5 of the secondary hydraulic servos 7c, 9c.
Then, both are controlled to be equal pressure, and further sent to the clutch modulation valve 79. The output hydraulic pressure from the clutch modulation valve 79 is selectively supplied to the low clutch C _L or the high clutch C _H except in the N and P ranges of (4). (1) D range-low mode The same hydraulic pressure is supplied to the first hydraulic chambers 55 and 56, the low clutch C _L is connected, and the hydraulic pressure is supplied to the second hydraulic chamber 59 on the secondary side in the upshift. Is
In the downshift, the hydraulic pressure is supplied to the primary-side second hydraulic chamber 57 only at the pulley ratio B or higher. That is, in the upshift, the manual valve 75 is operated to the D range position, the ports a and b, c and d, e and f are communicated, and the low / high control valve 76 is set to the low mode position.
The ports h and i, j and k, l and m communicate with each other, and the port g switches to communicate with the drain port Ex.
Is held.

Therefore, the hydraulic pressure from the clutch modulation valve 79 is applied to the low clutch C _L by the ports a and b of the manual valve 75, the ports h and i of the low / high control valve 76, and the port n of the low clutch relief valve 77. And o are supplied to the hydraulic servo for the roller clutch to engage the low clutch C _L. Further, the hydraulic pressure from the output port v of the primary regulator valve 71 is the same as that of the ratio control valve 72.
Is gradually increased to a hydraulic pressure corresponding to the target pulley ratio by the ports p and q and the manual valve 75.
2nd hydraulic chamber 5 on the secondary side via ports c and d of
9 is supplied. In this state, the high clutch C
_H is in an open state by communicating from the port g of the low / high control valve 76 to the drain port Ex, and the second hydraulic chamber 57 on the primary side includes the ports m and l of the low / high control valve 76 and the port of the manual valve 75. The ports f and e and the port s of the downshift relief valve 73 communicate with the drain port Ex. In the range in which the CVT 11 does not exceed the predetermined pulley ratio B due to the upshift, switching of the low / high control clutch 76 is mechanically restricted by the interlock rod 82.

As a result, the low clutch C _L is connected and the CVT 11 is connected to the first and second hydraulic chambers 56, 5
Secondary side hydraulic servo 9c where hydraulic pressure acts on both 9
The axial force due to becomes higher than the axial force due to the primary side hydraulic servo 7c in which the hydraulic pressure acts only on the first hydraulic chamber 55, the axial force is gradually increased, and the pulley ratio is increased.
At this time, the movable sheave 7b of the primary pulley 7 is U
Move to / D side. In this state, the engine torque transmitted from the input shaft 3 to the carrier 19c of the planetary gear 19 via the low clutch C _L and the constant speed transmission device 16 is C by the predetermined pulley ratio via the sun gear 19s.
It is taken out from the output gear 21 via the ring gear 19r while being regulated by the VT 11.

Regarding the downshift in the D range-low mode, in the region below the predetermined pulley ratio B, the ratio sensing valve 80 is passed through the sensor shoe 81.
However, in the state shown in FIG. 6, the hydraulic pressure from the output port v of the primary regulator valve 79 is stopped by the ratio sensing valve 80, whereby the second hydraulic chamber 57 on the primary side, which is necessary for downshifting, is supplied. Supply of hydraulic pressure becomes impossible. Even in this case,
By communicating the port q of the ratio control valve 72 with the drain port Ex, the oil pressure in the second hydraulic chamber 59 on the secondary side can be drained, so that downshifting to the neutral state is possible. on the other hand,
In the region where the pulley ratio B is equal to or more than the predetermined value, downshifting can be performed by the ratio sensing valve 80 and the like. That is, above the predetermined pulley ratio B, the movable pulley 7b of the primary pulley 7 moves to the U / D side, and the ratio sensing valve 80 moves downward in the figure via the sensor shoe 81. Therefore, the hydraulic pressure from the primary regulator valve 71 is guided to the downshift relief valve 73 via the check valve 85 as the ports t and u of the ratio sensing valve 80 communicate with each other. Therefore, by moving the downshift relief valve 73 upward in the figure to connect the ports r and s, the ports e and f of the manual valve 75 and the ports 1 and m of the low-high control valve 76 are connected. The hydraulic pressure can be supplied to the second hydraulic chamber 57 on the primary side. (2) D range-high mode Both primary and secondary side hydraulic chambers 55, 5
A hydraulic pressure equal to 6 is supplied, the high clutch C _H is connected, and in the upshift, the second clutch on the primary side is used.
Is supplied to the second hydraulic chamber 59 on the secondary side in the downshift. That is, in the upshift in the D range-high mode, the manual valve 75 is in the same D range position as in the previous low mode, but the low / high control valve 76 is switched to the high mode position, and the ports h and g, j and m, l and k communicate,
And the port i communicates with the drain port Ex.

Therefore, the primary regulator valve 7
The output hydraulic pressure from the output port v of No. 1 is the ports a and b of the manual valve 75, and the low / high control valve 7
6 is supplied to the high clutch hydraulic servo via the ports h and g to engage the clutch C _H , and the ports p and q of the ratio control valve 72, the ports c and d of the manual valve 75, and the low / high control. It is supplied to the second hydraulic chamber 57 on the primary side via the ports j and m of the valve 76. In this state, the low clutch hydraulic servo C _L is in a released state by being communicated with the drain port Ex from the port i of the low / high control valve 76, and the second hydraulic chamber 59 on the secondary side.
Communicate with the drain port Ex via the ports k and l of the low / high control valve 76, the ports f and e of the manual valve 75, and the port s of the downshift relief valve 73.

As a result, the high clutch C _H is connected and the CVT 11 is connected to the first and second hydraulic chambers 55, 5
Primary side hydraulic servo 7c where hydraulic pressure is supplied to 7
Is increased by the axial force of the secondary hydraulic servo 9c supplied only to the first hydraulic chamber 56, and in the axial force state corresponding to the torque transmission from the primary pulley 7 to the secondary pulley 9, the ratio is increased. By appropriately adjusting the control valve 75, the hydraulic pressure of the second hydraulic chamber 57 of the primary hydraulic servo 7c is adjusted,
The axial force of the primary pulley 7 is adjusted to obtain an appropriate pulley ratio (torque ratio). In this state, the torque transmitted from the engine to the input shaft 3 is
Is appropriately changed by the CVT 11 transmitted from the secondary gear 9 to the secondary pulley 9, and is further taken out from the output gear 21 via the high clutch C _H.

In the D range-high mode described above, CVT11 is smaller than the predetermined pulley ratio B (O
/ D side), the interlock rod 82 mechanically prohibits the low-high control valve 76 from being switched to the low mode. Also, C
Even when the VT pulley ratio is equal to or less than the predetermined pulley ratio B, unlike the case of the D range-low mode, the downshift is not prohibited. That is, D range-
Even in the high mode, since the hydraulic pressure from the output port v of the primary regulator valve 71 is stopped by the ratio sensing valve 80 in the state shown in FIG. 6, this hydraulic pressure is downshift relief valve 73, manual valve 75, low-high control valve 76. Is not supplied to the second hydraulic chamber 59 on the secondary side via. However, instead of this, the hydraulic pressure from the high clutch C _H
Check valve 85, downshift relief valve 73
The second hydraulic chamber 59 on the secondary side is supplied through the ports r and s, the ports e and f of the manual valve 75, and the ports l and k of the low-high control valve 76. As a result, in the D range-high mode,
Downshifting is possible in all areas of the pulley ratio. (3) R range In the R range, a predetermined hydraulic pressure is supplied to the first and second hydraulic chambers 55 and 57 of the primary hydraulic servo 7c, and at the same time, the first hydraulic chamber of the secondary hydraulic servo 9c. 56, and also to the low clutch hydraulic servo C _L. That is, in the R range, the manual valve 75 is in the R range position and the low / high control valve 76 is in the low mode position. Therefore, the hydraulic pressure from the output port v of the primary regulator valve 71 is applied to the ports a and b of the manual valve 75 and the port h of the low / high control valve 76.
And i to the low clutch hydraulic servo C _L , and the primary side via ports p and q of the ratio control valve 72, ports c and f of the manual valve 75, and ports 1 and m of the low high control valve 76. Is supplied to the second hydraulic chamber 57. Further, the port s of the downshift relief valve 73 is connected to the drain valve Ex.

As a result, the low clutch C _L is connected, and the CVT 11 is connected to the first and second hydraulic chambers 55, 5
Axial force by the primary side hydraulic servo 7c that applies hydraulic pressure to 7 becomes higher than that on the secondary side by the secondary side hydraulic chamber 9c due to only the first hydraulic chamber 56, and torque is transmitted from the primary pulley 7 to the secondary pulley 9. It becomes the corresponding axial force state and the ratio control valve 5
By adjusting 7, the hydraulic pressure of the second hydraulic chamber 57 of the primary hydraulic servo 7c is adjusted, and an appropriate pulley ratio is obtained. In this state, the pulley ratio of the CVT 11 is in a predetermined acceleration (O / D) state, and the engine torque from the input shaft 3 is transmitted to the carrier 19c of the planetary gear 19 via the low clutch C _L and the constant speed transmission device 16. The torque is transmitted from the primary pulley 7 to the secondary pulley 9 as well as transmitted to the sun gear 19s via the CVT 11, and these two torques are combined by the planetary gear 19 and extracted as reverse rotation to the output shaft 5 via the ring gear 19r. To be done. In the R range as well, as in the case of the predetermined pulley ratio B or less in the D range-low mode, supply of hydraulic pressure to the downshift relief valve 73 is prohibited by the sensor shoe 81 and the ratio sensing valve 80. Although downshifting is prohibited, in the R range, originally, engine braking is not particularly required. Therefore, even if downshifting is prohibited, no inconvenience occurs. (4) At the P range position and the N range position of the N, P range manual valve 75, both the low clutch C _L and the high clutch C _H are released, and both the primary side and secondary side hydraulic servos 7c are released. , 9c of the first hydraulic chambers 55, 56 are supplied with a predetermined hydraulic pressure. That is, in the manual valve 75, the ports c and d and the ports e and f communicate with each other, and the port b communicates with the drain port Ex. The low / high control valve 76 is held in the low mode position described above. The port q of the ratio control valve 72 communicates with the drain port Ex, and the ratio sensing valve 80 is held at the position shown in FIG. Therefore, the primary hydraulic servo 7c and the secondary hydraulic servo 9c both apply the same hydraulic pressure only to the first hydraulic chambers 55 and 56, and both the primary and secondary pulleys 7 and
9, the substantially equal axial force acts.

This hydraulic control mechanism is described in Japanese Patent Application No. 7-32.
No. 7663 (not yet published at the time of application).

Next, the control of the continuously variable transmission of this embodiment will be described.

FIG. 7 is a block diagram of the electronic control unit (ECU) 90. Reference numeral 91 is a sensor installed in the continuously variable transmission 1 for detecting the rotation speed of the input shaft 2 of the transmission, and 92 is a sensor. C
A sensor for detecting the rotation of the secondary pulley 9 of the VT 11, 93 is a vehicle speed sensor for detecting the rotation of the output shaft 5 of the continuously variable transmission, and 94 is a shift lever of the continuously variable transmission, that is, a manual valve is P, R, N. , D is a sensor for detecting where each shift position is located, 95 is a mode select switch for selecting a power mode based on the maximum power characteristic or an economy mode based on the best fuel consumption characteristic, and 96 is an accelerator (throttle) pedal. A sensor for detecting the throttle opening based on the above, 97 is a potentiometer installed in the engine, and a sensor for detecting the actual opening degree of the throttle, 98 is a switch for detecting that the brake pedal is in the depressed position, 99 Detects that your foot is off the accelerator pedal and the throttle is idle. Sensor 100 is a kick-down switch for detecting the state in which the accelerator pedal is depressed in full, 101, the oil temperature sensor of the transmission,
102 is a sensor that detects the engine speed.

The signals from the above-mentioned sensors are sent to the CP via the input processing circuit and the input interface circuit, respectively.
U, ROM or RAM. And the CPU
The control unit composed of, for example, a signal from the throttle opening sensor 96 according to the pulley ratio of the CVT 11 calculated based on the signals of the input rotation sensor 91 and the secondary shaft rotation sensor 92, and the depression amount of the accelerator pedal (accelerator operating means). Based on the above, the electronic throttle control means Q for outputting an electronic throttle opening signal to the electronic throttle system 109, and the CVT 11 is provided with a biasing force that deviates from the neutral position by a predetermined amount in accordance with the output torque by the electronic throttle control means Q. The hydraulic actuators 7c and 9c are provided with engine load control means R that acts on the hydraulic actuators 7c and 9c by a predetermined pressure difference ΔP. The electronic throttle control means sets the output torque of the engine within the limit torque set according to the pulley ratio of the belt type continuously variable transmission 11 at the time of starting from the neutral position where the rotation of the output gear described later becomes zero. The throttle opening is controlled so that the engine speed becomes a target rotation speed according to the operation amount of the accelerator pedal (operating means) that manually operates the engine within the throttle opening. Further, the electronic throttle control means R includes the hydraulic actuators 7c and 9c.
A predetermined pressure difference ΔP is applied to c, and the engine load is controlled so that the engine speed becomes the target speed.

On the other hand, on the output side, reference numeral 76c is a solenoid for the low / high control valve 76 for switching between the low mode and the high mode, which is turned on / off. Reference numeral 73a denotes a solenoid for the downshift relief valve 73 for draining the high-voltage side circuit, which is used during engine braking or neutral (N) described later.
It is operated during control and is composed of a duty or linear solenoid. Reference numeral 72a is a solenoid for the ratio control valve 72 for adjusting the hydraulic pressure for shift control, which is a duty or linear solenoid.
Reference numeral 77a denotes a solenoid valve for the low clutch relief valve 77, which is a duty solenoid.
Reference numeral 71a is a solenoid for the primary regulator valve 71 for controlling the line pressure, which is a linear solenoid. Then, each of the solenoids is driven via a solenoid drive circuit 106 that generates a predetermined voltage or output based on a signal from the output interface circuit, and the operation of each solenoid is checked by a monitor circuit 107 and a failure is detected. Is determined and self-determination is performed.

Reference numeral 109 is an electronic throttle system section for engine control, and 110 is a processing circuit for outputting a drive signal of the electronic throttle stepping motor and for inputting feedback information. 112
Is composed of an indicator lamp and the like, and the electronic control unit 90
Is a checker member that outputs the self-diagnosis result at the time of the failure, and 113 is a circuit for outputting the self-diagnosis result at the time of the failure. Reference numeral 115 is a display device for displaying the state of the continuously variable transmission such as a power mode and economy mode display lamp, and 116 is a drive circuit therefor. Further, the continuously variable transmission 1 is directly transmitted from the engine output shaft 2 to the input shaft 3 only via the damper device 12, and has a conventionally required starting device, that is, a torque converter, a fluid coupling, an electromagnetic powder clutch. Or no input clutch is required. Therefore, in the D and R ranges, it is necessary to perform N control in which the continuously variable transmission 1 automatically becomes neutral (neutral) when the vehicle is stopped.

The N control is operated by the neutral control means N which operates based on the judgment that the judgment means requires the neutral state, and the axial forces of the primary pulley 7 and the secondary pulley 9 are substantially equal. Control to be. That is, at least the difference between the axial forces of the primary and secondary pulleys is expressed by the difference between the axial forces of the two pulleys determined by the input torque of the CVT and the pulley ratio at that time when the output torque direction is positive. A range in which the magnitude relationship is not reversed from the difference between the axial force of the primary and secondary pulleys determined by the CVT input torque and the pulley ratio at that time when the output torque direction is negative or when the output torque direction is negative. Is controlled to be a small value. Specifically, in a state where the hydraulic pressure is supplied to both the first hydraulic chambers 55 and 56 in both the primary and secondary hydraulic actuators 7c and 9c, the second hydraulic chambers
The hydraulic pressure in the hydraulic chambers 57 and 59 is released to equalize the axial forces of the pulleys 7 and 9.

Regarding the N control, that is, the control in which the CVT 11 is self-converged so that the CVT 11 is in the neutral state by setting the difference between the axial forces of the primary side hydraulic actuator 7c and the secondary side hydraulic actuator 9c within a predetermined range. , Japanese Patent Application No. 7-66234 or Japanese Patent Application No. 7-12870
No. 1 is described in detail.

In the above-mentioned prior application, when the CVT shifts in the U / D (deceleration) direction due to the negative torque condition during coasting, and the engine speed becomes lower than the idle speed, the positive torque condition occurs. CVT is O / D (acceleration)
It is described that the CVT 11 self-converges to the gear neutral (GN) point by repeating the shifting in the direction according to the vehicle speed, and the vehicle is stopped and stably held at the neutral point. Based on the difference between the axial forces of the primary pulley and the secondary pulley in the positive torque state and the negative torque state, the CVT 11 is stably held at the GN point by making both axial forces substantially equal (within a predetermined range).
The self-convergence to the GN point during the N control has a slow convergence speed due to the switching between the negative torque and the positive torque, and in the present embodiment, the engine is controlled so that the N torque is always in the positive torque state. Control the torque.

N based on the input (engine) torque control
The principle of self-convergence during control will be described. The Ogasawara equation shown in the following equation 1 is an equation (balance equation) indicating the axial force of the V-belt.

[0058]

[Formula 1] Here, F _DV is the axial force of the drive side pulley (secondary pulley), F _DN is the axial force of the driven side pulley (primary pulley),
φ ₁ is the drive-side pulley belt winding angle, φ ₂ is the driven-side pulley belt winding angle, r ₁ is the drive-side pulley effective radius, r ₂
Is the effective radius of the driven pulley and T _in is the input (engine) torque.

The above equation 1 is simplified and summarized as equation 2.

[0060]

[Formula 2] Here, f (I _p ) is a function that changes according to the pulley ratio (r ₂ / r ₁ ).

That is, the driving-side axial force F _DV that balances the driven-side axial force F _DN requires a larger force as the input torque T _in increases. Here, when both axial forces are made substantially equal by the above N control, the relationship of the axial forces becomes the following Expression 3.

[0062]

[Formula 3] Here, the balance value (theoretical value) in Equation 3 is φ ₁ <
If it is controlled so that F _DV ≈ F _{DN in the} relation of φ _2, the axial force of the drive side (secondary) pulley becomes excessive, and this axial force difference ΔF _DV closes the drive side pulley and CVT becomes OVT. Converge in the / D direction. That is, the difference between the actual value of the axial force and the balance value (theoretical value) ΔF _DV [ΔF _DV = F
_DV (actual value) −F _DV (balance value)] converges to the GN point, and this load difference (ΔF _DV ) becomes more significant as the input torque (T _in ) becomes larger, as is clear from Equation 3. It will be a heavy load.

Input torque T based on the above Ogasawara equation 1
FIG. 8 shows the relationship between _in (= T _E ; engine output torque) and the convergence force ΔF to the GN point. From FIG. 8, it is clear that the convergence force ΔF increases as the input torque T _in increases at each pulley ratio (1.3, 1.0, 0.8, GN) of the CVT, and the convergence force ΔF increases. The convergence speed increases due to the increase in force.

The convergence mechanism of the N control by controlling the input torque, that is, the engine torque (when the throttle opening is fixed) will be described with reference to FIG. If the vehicle speed falls below the predetermined speed in the vehicle deceleration state in which the accelerator pedal is released or the brake pedal is further depressed, N
The output torque of the engine (the input torque of the CVT) is increased by the N control when the control is started, and the pulley ratio of the CVT self-converges toward the GN point at a high speed as described above. The engine torque gradually increases and the engine torque gradually decreases (start → A).

Then, the convergence speed becomes slow due to the decrease in the engine torque, and the increase in the engine speed is also suppressed. On the other hand, as the vehicle continues to decelerate, the above C
When the convergence of VT cannot follow the deceleration of the vehicle, the engine speed decreases (A → B). When the engine speed decreases, the engine torque increases, so that the convergence speed increases again as described above, and it becomes possible to follow the deceleration of the vehicle (after B). After that, repeat this to CV
The pulley ratio of T converges to the GN point and the vehicle is stopped.

Next, the principle of generation of the creep torque of the automatic continuously variable transmission (IVT) 1 which is in the neutral state at the GN point of the CVT described above will be described with reference to FIG. An automatic transmission (A / T) having a multi-stage transmission mechanism such as a conventional planetary gear includes a torque converter (starting device) between an engine output shaft and a multi-stage transmission mechanism input shaft. When the transmission mechanism input shaft is stopped, a torque increased according to the engine speed is applied to the input shaft (stall torque) to smoothly start the vehicle. The automatic continuously variable transmission 1 is automatically held in the neutral position by the N control, and in the neutral state, generates the forward torque (creep torque) similar to that of the torque converter.

As described above, the force F _N that causes the CVT to self-converge to the GN point from the forward drive region or the reverse drive region is generated, and the achievement of the GN point causes the CVT to be unloaded or almost unloaded. It becomes a state. On the other hand, the CVT 11 itself is in a stable state in which the primary and secondary pulleys are counteracted by the belt tension, that is, the pulley ratio becomes 1.0, and the force F _A is generated toward the pulley ratio 1.0. Therefore, when the CVT becomes the GN point and the automatic continuously variable transmission (IVT) 1 becomes in the no-load state, at the same time, the CV becomes
_A force F _A is generated toward the point where the pulley ratio is 1.0, which is the stable convergence point of T itself. The force F _A toward the pulley ratio of 1.0 in the wireless load conditions, the force F _A force CVT is toward the GN point under load by departing from GN point by F _N is, as shown in the enlarged model diagram A vortex state is generated in the forward direction, which causes forward creep torque.

CVT is set to O by setting an area difference between the first hydraulic chambers 55 and 56 on the primary and secondary pulley sides so as to oppose the force F _A toward the pulley ratio of 1.0.
When the axial force F _O in the / D direction is applied, the force F _A is canceled and the forward creep torque disappears. The axial forces F _A and F _O are not affected by the input torque and the transmission efficiency, and the axial forces of the primary and secondary pulleys during N control are biased by a predetermined amount from the equal value,
Any creep torque can be obtained.

FIG. 11 is a diagram showing the relationship between the axial torque difference (upper stage) and the supplied hydraulic pressure difference (lower stage) between the primary and secondary pulleys and the creep torque. The upper stage of the horizontal axis shows the secondary pulley with respect to the primary pulley. The axial force difference is shown, and the vertical axis shows the torque amplification ratio in the forward direction and the downward direction in the reverse direction on the vertical axis. As shown in FIG. 1, the first hydraulic chamber 5 in both the primary and secondary pulleys 7 and 9
When the effective pressure receiving areas of 5,56 are equal (A _P = A _S ),
The axial force difference between both pulleys becomes zero, and creep torque having a predetermined torque amplification ratio is generated in the forward direction. The lower part of the horizontal axis is set so that the effective torque receiving areas of the primary side and secondary side hydraulic actuators are made different (A _P > A _S ), and the creep torque becomes approximately 0 when equal hydraulic pressures are supplied to both hydraulic actuators. In the automatic continuously variable transmission described above, the difference between the hydraulic pressure supplied to the secondary hydraulic actuator and the hydraulic pressure supplied to the primary hydraulic actuator is shown. In this case, when it is ensured that the automatic transmission 1 is in the forward drive state and the creep torque in the forward drive direction based on the torque similar to that of the conventional torque converter is set, the differential pressure in the secondary side hydraulic chamber is higher than that in the primary side. A pressure P _C is supplied. In the forward travel ensuring pressure / hydraulic pressure range, a predetermined forward creep torque is generated even during the N control, and the vehicle creeps in the forward direction when the brake is released.

If an area difference is provided in the hydraulic chambers of both pulleys (A _P > A _S ) and F _A = F _O is set, the creep torque becomes zero. However, in actual control,
There is an error in the hydraulic pressure that is set, and within this error range, torque may be generated in the reverse direction even if the driver is aware of the forward state, so a slight differential pressure is supplied to the secondary side to make the vehicle The range P _min is set on the forward side in the range where the vehicle does not move forward (less than the required torque for moving the vehicle). In the range of P _min , in N control, the vehicle remains stopped even if the brake pedal is released.

Further, when a hydraulic pressure higher than that on the secondary side is supplied to the hydraulic chamber of the primary pulley, the hydraulic pressure error is within the range in which the automatic continuously variable transmission is guaranteed to be in the reverse drive state. In the reverse drive guaranteed pressure / hydraulic range, reverse creep torque is generated even in the N control state, and the vehicle creeps in the reverse direction when the brake pedal is released.

Next, with reference to FIG. 12, an embodiment will be described in which the creep torque becomes substantially 0 when the same hydraulic pressure is supplied to the first hydraulic chambers of both the primary and secondary pulleys. 12, the primary pulley side hydraulic actuator 7c is the same as that of FIG. 1, but the secondary pulley side hydraulic actuator 9d is slightly different. The effective pressure receiving area of the second hydraulic chamber 59 of the secondary pulley side hydraulic actuator 9d is the same as that of the second hydraulic chamber 57 on the primary side, but the effective pressure receiving area (A _S ) of the first hydraulic chamber 56 is The area is set smaller than the area (A _P ) of the first hydraulic chamber 55 on the primary side by a predetermined amount (A _P > A _S ). Therefore, in the N control in which the hydraulic pressures of the second hydraulic chambers 57 and 59 on the primary and secondary sides are released and the same predetermined hydraulic pressure is supplied to both the first hydraulic chambers 55 and 56, the primary pulley 7
Axial force F _S of the secondary pulley 9 to the axial force F _P of a predetermined amount smaller. The difference between the axial forces of these pulleys (F _P -F
_S ) becomes the force F _O that opposes the force F _A toward the pulley ratio of 1.0, which cancels the forward creep torque. In this state, in N control, the creep torque becomes substantially 0, and even if the brake is released,
The vehicle is held stationary.

Further, a partially modified continuously variable transmission and hydraulic control mechanism will be described with reference to FIGS. 13 and 14.
The continuously variable transmission 1 is the same as the previous embodiment except that the hydraulic actuator is different, and therefore, the same portions are denoted by the same reference numerals and the description thereof will be omitted.

Primary and secondary pulleys 7, 9
Flange members 120 and 121 are fixed to the ends of the bosses of the fixed sheaves 7a and 9a, respectively, and the drum members 122 and 1 are attached to the rear surfaces of the movable sheaves 7b and 9b.
23 is fixed. Hydraulic chambers 125 and 126 are formed between the rear surfaces of the movable sheaves 7b and 9b and the flange members 120 and 121, respectively, and single chamber type hydraulic actuators 7e and 9e are formed on the primary side and the secondary side, respectively. Oil pressure is supplied to the oil pressure chambers 125 and 126 via oil passages 127 and 129, respectively, and preload springs 130 and 131 are contracted.

The hydraulic actuators 7e and 9e on the primary and secondary sides are connected to the hydraulic chamber 12 on the primary side.
Effective pressure receiving area A _P 5 is set a predetermined amount larger than the effective pressure receiving area A _S of the secondary side of the hydraulic chamber 126 (A
_P > A _S ), and when the hydraulic pressure equal to both hydraulic chambers 125 and 126 is supplied, the creep torque is set to be substantially zero.

FIG. 14 shows a hydraulic control mechanism applied to the hydraulic actuators 7e, 9e consisting of the single chamber, and is different from that shown in FIG. 6 in that there are two regulator valves. This hydraulic control mechanism
It has a high regulator valve 71 ₁ and a low regulator valve 71 ₂ .

In the D range low mode, the hydraulic pressure regulated by the high regulator valve 71 ₁ is the clutch modulation valve 79 and the manual valve 75.
Is supplied to the low-clutch hydraulic servo C _L via the ports a and b, the low-high control valve 76 ports h and i, and the low-clutch relief valve 77.
The pressure is adjusted by the ratio control valve 72, and the adjusted hydraulic pressure is supplied to the secondary side hydraulic chamber 126 via the ports c and d of the manual valve 75 and the ports j and k of the low-high control valve 76. On the other hand, the hydraulic pressure from the low regulator valve 71 ₂ is supplied to the manual valve 7
5 is supplied to the primary-side hydraulic chamber 125 via the ports e and f and the ports 1 and m of the low-high control valve 76.

As a result, the low regulator valve 71
A constant pressure consisting of low pressure based on ₂ is the primary side hydraulic chamber 1
25, the high regulator valve 71 ₁
Is supplied to the secondary-side hydraulic chamber 126,
Further, the CVT 11 is shifted by appropriately adjusting the high pressure by the ratio control valve 72.

Then, in the N control, the ratio control valve 72 is switched to the down state, and the constant low pressure from the low regulator valve 71 ₂ causes the ratio sensing valve 80, the downshift valve 73, and the port x of the ratio control valve 72. , Y, the ports c and d of the manual valve 75, and the ports j and k of the low-high control valve 76 are supplied to the secondary hydraulic servo 126.

[0080] Thus, the primary-side hydraulic chamber 125 and the secondary-side hydraulic chamber 126 hydraulic action equal from low regulator valve 71 _2, creep based on area difference of the hydraulic chambers (A _P> A _S) Torque is almost 0
Becomes

In the D range high mode, the low / high control valve 76 is switched, the hydraulic pressure from the clutch modulator valve 79 is supplied to the high clutch hydraulic servo C _H , and the constant pressure from the low regulator valve 71 ₂ is maintained. The low pressure is supplied to the secondary hydraulic chamber 126, the high pressure from the high regulator valve 71 ₁ is supplied to the primary hydraulic chamber 125, and the high pressure is appropriately adjusted by the ratio control valve 72.

In the hydraulic actuator shown in FIG. 13, the hydraulic torque chambers are set so that the creep torque is substantially zero by providing an area difference between the hydraulic chambers as in the case shown in FIG. It is needless to say that the creep torque in the forward direction may be generated by supplying the same hydraulic pressure to both hydraulic chambers, as in the case shown in FIG. 1, without providing the area difference.

Next, the control of the present embodiment will be described with reference to the main flow chart shown in FIG.

S1 is a step for initializing all conditions at the start of the program, and S2 is T / M.
This is a step of calculating the current T / M input shaft (primary shaft), secondary shaft, output shaft rotation speed from the signals of the input (primary shaft) rotation sensor 91, the secondary shaft rotation sensor 92, and the vehicle speed sensor 93, and S3 is Is a step of calculating the pulley ratio I _P (primary shaft rotational speed / secondary shaft rotational speed) from the input shaft and the secondary shaft rotational speed. Further, S4 is a step of calculating the current throttle opening (θd) of the driver from the signal of the throttle opening sensor 96 of the driver, and S5 is T / M.
This is a step of calculating the current oil temperature of the transmission (ATF oil temperature) from the signal of the temperature sensor 101. Then, S6 is the Lo-Hi mode switching solenoid 76.
Based on the signal of c, it is a step of judging whether the current state of the continuously variable transmission is in the low mode, S7 is a step of calculating the target throttle opening degree (θ ^* ) in the low mode, and S8 is a step. This is a step of comparing the current vehicle speed V by the vehicle speed sensor 93 with a predetermined minimum vehicle speed V _min to determine whether or not the vehicle is in a stopped state.

Then, in S9, when the current pulley ratio is the pulley ratio B shown in FIG. 4 (for example, 1.3), that is, in the coast state, smooth torque transmission is hindered by the transmission efficiency and large engine braking is performed by the large gear ratio. Is a step of determining whether or not the pulley ratio is less than the pulley ratio at which there is a possibility that the operation (transmission failure due to the coast state) may occur. The estimated engine torque value Te that is set is
_This is a step of determining whether it is less than _min , and S
Reference numeral 11 is a coast avoidance control subroutine described later for preventing the engine from being negatively driven (coast).

Further, S12 is a step for performing the shift control in the low mode described above. However, when the forced off-up flag determined in step S11-4 described later is turned on, the shift speed is set to the maximum and upshifting is performed. S13 is a step of outputting a pressure signal to the hydraulic control solenoid 71a to control the calculated line pressure by the linear solenoid pressure, and S14 is a step of outputting a signal to the ratio control solenoid 77a to shift gears. Is a step for controlling the shift pressure of the solenoid by the solenoid pressure.
Reference numeral 15 is a step of outputting a signal to the Lo-Hi mode switching solenoid 76c to control the clutch switching between the low mode and the high mode by the solenoid.
Outputs a signal to the electronic throttle system 109,
In this step, a target throttle opening signal is output to control the opening of the electronic throttle.

Further, S17 is a start-up control subroutine from after-mentioned stop, and S18 is a step for calculating a target throttle opening degree θ ^* in the high mode,
Reference numeral 19 is a step for performing the shift control in the high mode described above.

Next, the coast avoidance control, which is a characteristic part of the present invention, will be described with reference to the flowchart shown in FIG. 16 and the time charts shown in FIGS. 17, 18, and 19. That is, the vehicle speed V is higher than the minimum speed V _min ,
The vehicle is in a predetermined traveling state, and the pulley ratio I _P is within a predetermined region where transmission failure due to the coast state may occur (I _P <B), and when the accelerator is off, the engine torque Te is lower than the predetermined minimum engine torque Te _min . If it is smaller, the coast avoidance control is activated. First, from the current engine speed, the target throttle opening θ ^* by the electronic throttle system 109 is calculated from the map shown in FIG. 20, and is reset to the target throttle opening set in step S7 (S11- 1). Then, it is determined by the brake switch 98 whether the brake is operated.

When the brake is on, the above-mentioned N control is activated (S11-5). FIG. 17 is a time chart of gear change and engine control when the brake is turned on. First, C
When the VT 11 is at the GN point (eg, 0.692) and the engine is idling and the vehicle is stopped, depressing the throttle pedal performs the start control described later, and the electronic throttle system 109 controls the engine throttle opening. Engine output torque increases,
As a result, the above-mentioned forward creep torque is generated to move the vehicle forward, and the CVT 11 shifts in the U / D direction and accelerates. When the pulley ratio is within a predetermined range ( _IP <B) due to traffic congestion or the like, and the accelerator pedal is released and the brake pedal is depressed, the neutral (N) control starts.

In the N control, the hydraulic pressure in the secondary side second hydraulic chamber 59 shown in FIGS. 1 and 12 is released, and
Low pressure from the low regulator valve 71 ₂ is supplied to the secondary side hydraulic chamber 126 shown in FIG.
The same hydraulic pressure is supplied to. At the same time, the engine throttle opening θ ^* is controlled by the electronic throttle system, the engine outputs a predetermined low torque T _{E min} , and the CVT1
The torque is input to 1. Thereby, as described above, the CVT 11 converges toward the GN point at a relatively high speed.

At this time, the input rotation sensor 91 detects the number of rotations of the input shaft, and the electronic control system controls the throttle opening θ ^* based on the map shown in FIG. That is, the coast state in which torque is transmitted from the wheels toward the engine is prevented, and the above N
The control is controlled so that a predetermined engine torque is output so that the control convergence force does not have a negative value.

Then, in response to the decrease in vehicle speed, CVT1
1 shifts in the O / D (downshift) direction, the throttle opening θ ^* also gradually decreases, and when the engine speed reaches a set engine speed Ne that outputs a constant low torque T _{E min} , the throttle The opening degree θ ^* is held at a position where the set rotational speed N _E and the predetermined torque T _{E min} are obtained (θ ^{* is} constant). With the predetermined torque T _{E min} maintained, the CVT 11 automatically converges toward the GN point,
The vehicle stops at the point N.

Next, when the brake pedal is not depressed and is in the off state, the N control flag set in the N control (S11-5) is reset (S11-3),
Then, the compulsory off-up flag is turned on (S11-
4). This state will be described with reference to the time charts shown in FIGS. The start control until the accelerator is turned off will be described later and is the same as that shown in FIG.

As shown in FIG. 18, when the accelerator is turned off, a signal is output to the ratio controlling solenoid 72a,
The ratio solenoid valve 72 is in the maximum output state and is supplied to the second hydraulic chamber 59 on the secondary side (FIGS. 1 and 12) or the secondary hydraulic chamber 126 (FIG. 13). As a result, the CVT 11 is shifted in the U / D (upshift) direction at the highest speed, and the predetermined pulley ratio B (blocking smooth torque transmission due to transmission efficiency and large engine braking action due to large gear ratio (coast state) It is quickly escaped from the area where (transmission failure due to) may occur.

At this time, the engine throttle opening θ ^* is controlled by the electronic control system so that the engine output is constant torque T _E.
It is controlled to output _min . In this state, the accelerator is off and the brake is off, and the vehicle is running by inertia, but the upshift of the CVT 11 causes the rotation speed of the input shaft 3 to rapidly decrease, and the throttle opening accordingly. The degree θ ^* also decreases. And CVT1
When 1 becomes equal to or higher than the pulley ratio B which may cause a transmission failure due to the coast state, the coast avoidance control is canceled and the normal low mode shift control (S12) is performed, and at the same time, the throttle opening becomes idle. .

On the other hand, as shown in FIG. 19, in the coast avoidance control when the brake is off, the throttle opening degree is changed to the engine minimum rotation speed Ne as the rotation speed of the input shaft 3 is decreased by the fastest upshift of the CVTll. _min , that is, the engine speed will be stopped if it decreases more (eg 6
00 rpm), the throttle opening is maintained in a constant state so that the engine output torque is maintained at a constant low torque T _{E min} and the engine speed is maintained at the minimum engine speed Nemin. At the same time, the CVTll is controlled by the low-mode shift control so as to gradually shift in the upshift direction by an amount commensurate with the deceleration accompanying the coasting of the vehicle, and the pulley ratio of the CVT may cause a transmission failure due to the coast state. When the pulley ratio B is exceeded,
The throttle opening becomes idle.

The description of the coast avoidance control is given in D
Although the case where the vehicle is in the range and is in the forward drive state has been described, even when the vehicle is in the reverse drive state with the R range, the predetermined value on the O / D side from the GN point of the CVTll,
There is a value that may cause a smooth transmission hindrance due to the transmission efficiency and a large engine braking action due to a large gear ratio (transmission failure due to the coast state), and the pulley ratio within the predetermined value and the accelerator pedal When the dull is not operated or the engine output torque is T _{E min} or less, the coast avoidance control is similarly performed.

Next, the flowchart of FIG. 21 and FIG.
Step S17 (see FIG. 1)
The start control of 5) will be described. First, since the vehicle speed V has fallen below the set minimum speed V _min , the N control flag is reset (S17-1), and the idle switch 99
Determines whether the throttle opening is in the idle state. When in the idle state, the neutral (N) control is performed and the brake pedal switch 98 is further activated.
Determines whether the brake is on (S17-1).
0).

As shown in FIG. 22, when the vehicle is stopped, CV
The pulley ratio I _{P of} Tll is at the GN point and the brake B is
_rk is ON (actuated), and throttle opening θ
Is in the idle state, the hydraulic pressure difference ΔP acting on both the primary and secondary pulleys is at the minimum value P _min that does not move the vehicle (S17-11), and the target engine speed N _e ^* is It is at a predetermined minimum value (idle speed). From this state, as shown at point A, when the brake pedal B _rk is released to become the OFF (release) state,
The second hydraulic chamber 59 (see FIG. 1) on the secondary pulley (drive) side.
12) Further, a predetermined hydraulic pressure is supplied to the hydraulic chamber 126 (FIG. 13), and the hydraulic pressure difference ΔP acting on both pulleys becomes the above-mentioned Pc, and the predetermined creep is caused by the axial force difference acting on both pulleys based on the hydraulic pressure difference. A force is generated (S17-12).

As shown at point B, the driver intends to start the vehicle by depressing the accelerator (throttle) pedal. However, as will be described later, the torque converter characteristics of a normal automatic transmission (A / T) are Similarly, a predetermined target engine speed N _ei ^{* of} the engine is set (S17-3) so that a characteristic that the stall torque increases according to the amount of depression of the throttle pedal is set (S17-3), and the target engine speed is reached. , The throttle opening θ ^* is reset (S17-4). At this time, in the automatic continuously variable transmission 1, a large torque is generated due to the torque circulation in the vicinity of the GN point. Therefore, it is necessary to significantly restrict the engine output torque by limiting the allowable torque of the belt or the like. Therefore, the actual throttle opening degree θ ^* is maintained at a low opening degree in response to a command from the electronic throttle control means Q from the control unit with respect to the depression amount θd of the accelerator pedal (accelerator operation means). Then, the hydraulic pressure to the second hydraulic chamber (FIGS. 1 and 12) or the hydraulic chamber 126 (FIG. 13) on the secondary pulley side is increased, and the differential pressure ΔP of both pulleys is set to the right in FIG. 11. Is set (S17-
5).

In the starting control, the CVT 11 is near the gear neutral (GN) position and the output torque of the continuously variable transmission 1 is greatly amplified as shown in FIG. 24, the input torque of the continuously variable transmission, that is, the output torque Te of the engine is restricted in accordance with the pulley ratio, as shown in FIG. That is, when the pulley ratio I _P is larger than 0.9, the above output torque is not greatly amplified and the above restriction is not required, but it is significantly restricted in the vicinity of the GN point at the time of starting, and the electronic throttle control means Q The maximum value θ ^* _max of the target throttle opening is limited to 2-3%, for example. In this way, as shown in FIG. 24, the maximum throttle opening θ ^* _max is determined from the pulley ratio.

Further, as shown in FIG. 24, the driver controls acceleration and the like by the depression amount (operation amount) of the accelerator pedal, and the accelerator pedal operation amount θd is 100.
% Of the above-mentioned maximum target throttle opening θ ^* _max to the full throttle (100%) of the driver operation throttle opening θd (for example, 50%).
%, 20%), the target throttle opening degree θ ^{* at} each pulley ratio is set. That is, as shown in FIG.
The target throttle opening θ ^* is set by multiplying the driver operation throttle opening θd with the maximum target throttle opening θ ^* _max based on the pulley ratio. Further, the maximum torque on the set target throttle opening θ ^* is determined as the engine output torque Te.

As described above, when the throttle opening is extremely low, the equal throttle line of the engine output characteristic consisting of engine torque-rotational speed with the throttle opening as a parameter is sloping downward. That is, when the throttle opening is kept constant at an extremely low opening, the engine torque tends to decrease as the engine speed increases.

In a normal running state (low mode, high mode), the engine speed fluctuates due to a load change such as a driver's brake operation or a sudden change in running resistance.
It is necessary to consider the maximum torque on the equal throttle line at the corresponding target throttle opening θ ^*, but in the stall state of the start control, there is no disturbance effect such as the above-mentioned running resistance change and no engine load. It is determined by the control of the transmission. Therefore, the engine target speed N _ei ^* at the time of start (stall) based on the driver operation throttle opening θd is obtained from the map shown in FIG. 23 (S17-3), and the engine target speed N _ei ^* and the above From the set engine output torque Te, the target throttle opening θ ^* is updated (reset S17-4) by the map of the engine output characteristic, and the CVT pulleys are set according to the determined engine output torque Te. Is set (S17-5), and the engine load is controlled so that the engine speed becomes the target value (engine load control means R).

Then, as shown in FIG. 21, when the current engine speed N _e is larger than the value obtained by adding a predetermined value d to the target speed Nei ^* for start control (N _e > N _ei ^* +)
d) (S17-6), control is performed so that the differential pressure ΔP, which is large on the secondary side, becomes larger, and CVT11
Is urged in the U / D direction (actually at the GN point) to increase the engine load and lower the engine speed Ne (S17-7). If N _e > N _ei ^* + d is not satisfied, the current engine speed Ne is compared with a value obtained by subtracting the predetermined value d from the start control target speed N _ei ^* (S17-
8), if N _e > N _ei −d, that is, if the current engine speed N _e is within the target speed range, it is maintained as it is,
When N _e <N _ei −d, the differential pressure ΔP that is large on the secondary side is controlled to be small, and C
The VT11 biasing the O / D direction (actually in GN point) decreases the load of the engine by increasing the engine speed N _e (S17-9). That is, the engine load control means R controls the pressure difference between the two hydraulic actuators so that the engine speed becomes the target speed.
Change the load on the engine.

By controlling the throttle opening by the electronic throttle control means Q and controlling the engine load by the engine load control means R, the engine output torque Te and the engine speed N _ei can be set arbitrarily. When the accelerator pedal is stepped on, the engine speed increases and the input torque of the transmission increases accordingly, and the same characteristics as those of the conventional torque converter can be obtained. Since the engine output torque and the engine speed can be arbitrarily set, not only the characteristics of the torque converter but also the characteristics of another starting device such as an electromagnetic clutch or a manual clutch can be freely set.

The vehicle starts based on the set engine speed, engine output characteristic (stall characteristic) and creep torque. As a result, as shown in FIG. 22, when the pulley ratio I _{P of the} CVT 11 is slightly changed in the U / D direction and the vehicle speed reaches the predetermined low speed value V _min , the above-described start control is performed as indicated by point C. Is shifted to the normal shift control, and the engine speed is set to the target speed N _e ^* at the normal time. That is, as shown in FIG. 23, in starting control for generating the stall torque, the engine speed N _e is the throttle opening θd as shown by N _ei ^*.
The engine speed N _e is controlled based on the maximum output characteristic N _ep ^* or the best fuel consumption characteristic N _ee ^* in the normal shift control. In this state, CVT1
The pulley ratio of 1 is shifted in the U / D direction, and the vehicle speed V is increased.

Then, as shown at the point D, when the pulley ratio of the CVT becomes about 1.0, the electronic throttle control means for preventing the excessive torque from acting on the continuously variable transmission such as the belt described above. The control of the throttle opening θ ^* due to is released, and it corresponds to the amount of depression θd of the throttle pedal by the driver.

In the above description, the shift lever is in the D range and is started in the forward direction. However, when the shift lever is in the R range and is started in the reverse direction, the second shift on the primary pulley side is performed. A predetermined oil pressure is supplied to the oil pressure chamber 57 (FIGS. 1 and 12) or the oil pressure chamber 125 (FIG. 13), which is within the backward movement guarantee hydraulic pressure error range in FIG. 11.
A similar creep force is generated in the reverse direction. Except for this point, the operation is the same as that shown in FIG.

Although the above-described embodiment uses the belt type continuously variable transmission, the present invention is not limited to this, and can be similarly applied to those using other continuously variable transmissions such as a toroidal type.

[Brief description of drawings]

FIG. 1 is a front sectional view showing a continuously variable transmission according to the present invention.

FIG. 2 is a velocity diagram thereof.

FIG. 3 is a diagram showing an engaged state of each clutch.

FIG. 4 is a diagram showing a change in output torque with respect to a torque ratio of the belt type continuously variable transmission (CVT).

FIG. 5 is a diagram showing a change in output rotation speed with respect to the CVT torque ratio.

FIG. 6 is a diagram showing a hydraulic control mechanism applicable to the continuously variable transmission according to the present invention.

FIG. 7 is a block diagram showing the electric control mechanism.

FIG. 8 is a diagram showing a relationship between an input torque and a convergence force at each pulley ratio.

FIG. 9 is a diagram for explaining a mechanism of convergence to a GN point during N control by input torque control.

FIG. 10 is a diagram for explaining a mechanism for generating creep torque.

FIG. 11 is a diagram showing the relationship between the torque amplification ratio during creep, the axial force difference between both pulleys, and the hydraulic pressure differential.

FIG. 12 is a front sectional view showing a continuously variable transmission in which the effective pressure receiving area is changed.

FIG. 13 is a front sectional view showing a continuously variable transmission according to another embodiment.

FIG. 14 is a diagram showing the hydraulic control mechanism.

FIG. 15 is a main flowchart showing control of the continuously variable transmission according to the present invention.

FIG. 16 is a flowchart showing a subroutine of coast avoidance control.

FIG. 17 is a time chart at the time of braking on the coast avoidance control.

FIG. 18 is a time chart when the brake is off in the coast avoidance control.

FIG. 19 is a different time chart at the time of braking off the coast avoidance control.

FIG. 20 is a diagram showing a map of an electronic throttle target opening degree with respect to an engine speed.

FIG. 21 is a flowchart showing a start control subroutine.

FIG. 22 is a time chart thereof.

FIG. 23 is a diagram showing a target engine speed in each state.

FIG. 24 is a diagram showing an input torque limit at each pulley ratio.

FIG. 25 is a diagram showing a process of determining a target throttle opening θ ^* at the time of stall.

[Explanation of symbols]

1 continuously variable transmission 2 power source (engine) output shaft (crankshaft) 3 input shaft (first shaft) 5 axle (third shaft, front axle) 6 output shaft (fourth shaft, counter shaft) 7 first Rotating member (primary pulley) 7c, 7e Shift operating means (hydraulic actuator) 9 Second rotating member (secondary pulley) 9c, 9d, 9e Shift operating means (hydraulic actuator) 10 Belt 16 Constant speed transmission device 19 (simple) planetary gear 19c 1st rotary element (carrier) 19s 2nd rotary element (sun gear) 19r 3rd rotary element (ring gear) 55, 57, 125 1st (primary) pulley hydraulic chamber 56, 59, 126 2nd (secondary) pressure chamber C _L clutch (low clutch) of the pulley 90 the control unit 91 to 102 detecting means 109 electronic throttle system Q electrostatic Throttle control means R engine load control means

─────────────────────────────────────────────────── ─── Continued Front Page (51) Int.Cl. ⁷ Identification Code FI F02D 41/14 320 F02D 41/14 320A F16H 9/00 F16H 9/00 E 37/02 37/02 R 61/04 61/04 // F16H 59:18 59:18 (72) Inventor Shiro Sakakibara 10 Takane, Fujii-cho, Anjo-shi, Aichi Prefecture Aisin AW Co., Ltd. (72) Inventor Takeshi Inuzuka 10 Takane, Fujii-cho, Aichi-shi Aisin AW Co., Ltd. (72) Inventor Masashi Hattori 10 Takane, Fujii-cho, Anjo City, Aichi Prefecture Aisin AW Co., Ltd. (56) Reference JP-A-8-261303 (JP, A) JP 10-115357 (JP, A) JP 1-215635 (JP, A) JP 64-44346 (JP, A) JP 6-249329 (JP, A) JP 9-166215 (JP, A) Flat 6-331000 (JP, A) (58) Fields investigated (Int.Cl. ⁷ , DB name) F02D 29/00-29/06 B60K 41/00-41/28 F16H

Claims (4)

(57) [Claims] 1. An input shaft that interlocks with an output shaft of an engine, an output member that interlocks with wheels, a first rotating member, a second rotating member that interlocks with the input shaft, and both the first and second rotating members. A continuously variable transmission having a speed change operation means for changing the rotation ratio of the rotating member; and at least first, second, and third rotating elements, and based on the change of the rotation ratio of the continuously variable transmission, The first rotating element is used as the input shaft and the second rotating element is used as the second rotating element so that the torque transmission direction is changed between the rotating members and the output torque direction of the output member is changed. In a continuously variable transmission comprising: a rotating member, a planetary gear in which the third rotating element is interlocked with the output member, and a start-up from a neutral position where the rotation of the output member is 0, Output of the engine When the torque is within the throttle opening that is within the limit torque set according to the rotation ratio of the continuously variable transmission, the rotation speed of the engine depends on the operation amount of the accelerator operation means for manually operating the engine. An electronic throttle control means for controlling the throttle opening so as to achieve a target rotation speed, and a continuously variable transmission that is displaced from the neutral position by a predetermined amount according to the output torque of the engine by the electronic throttle control means. A continuously variable transmission, comprising: an engine load control unit that controls the shift operation unit so that an urging force acts. Wherein said engine load control means is formed by the rotational speed of the engine controls the shift operation means so that the target speed, the continuously variable transmission according to claim 1. 3. The electronic throttle control means controls the amount of operation of the accelerator operation means with respect to the limit torque based on the throttle opening degree by the electronic throttle control means when the accelerator operation means is operated to the maximum extent. formed by controlling the throttle opening degree such that the set output torque in proportion to claim 1 or 2 continuously variable transmission according. 4. A belt type continuously variable transmission having first and second pulleys, a belt wound around these pulleys, and a hydraulic actuator for exerting an axial force on the both pulleys. consists device, the engine load control means operates the hydraulic actuator, the belt type continuously variable transmission, the rotational speed of the engine is controlled to be the target speed, claim 2, wherein Continuously variable transmission.