JPS6125936B2 - - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION The present invention provides an automatic transmission for a vehicle in which a plurality of drive ratios are achieved between input and output shafts by selectively engaging a plurality of hydraulically operated friction engagement devices.
This invention relates to an improvement in a friction engagement device in which a band brake is used.

In general, the torque capacity of each frictional engagement device used in automatic transmissions is set according to the size of the engine output, and from the standpoint of space and cost, it is difficult to construct them with a large capacity. Can not.

Therefore, when one automatic transmission is used in various engines with different engine outputs, this is usually done by changing the oil pressure that operates the frictional engagement device according to the engine output of each engine. Due to the above-mentioned limitation of the supported torque capacity and also the limitations of the hydraulic control and hydraulic supply system, there is a limit to what can be done only by adjusting the working hydraulic pressure.

That is, the range of working oil pressure used in normal automatic transmissions for vehicles is 6 to 9 kg/ ^cm2 ,
If it is lower than this, the reliability of hydraulic control mechanisms such as hydraulic switching valves will decrease, and if it is higher than that, transmission efficiency will decrease due to increased oil pump loss and frictional drag loss at oil seals, etc. There is a problem that causes a decrease in fuel efficiency.

For this reason, in a friction engagement device employing a clutch disk, the receiving torque capacity can be increased by increasing the number of disks, but in a band brake, the only option is to increase the hydraulic servo area.

In other words, in the automatic transmission that has already been completed,
Since the band brake is compactly housed adjacent to various devices in the automatic transmission, increasing the band radius would require major equipment changes and
There is a drawback that it requires a great deal of effort to improve the band brake in order to obtain a band brake with desired characteristics, and there are also limits to the use of friction elements with a large friction coefficient μ.

By the way, in order to achieve a smooth shift by creating an overlap between the release of the band brake and the engagement of other frictional engagement devices during gear shifting, the hydraulic pressure acting on the release side pressure receiving surface of the hydraulic piston of the band brake and the above-mentioned The hydraulic pressure acting on the pressure receiving surfaces of other frictional engagement devices is controlled by the same hydraulic pressure, and the hydraulic pressure acting on both pressure receiving surfaces is maintained at a set low oil pressure by the accumulator effect by the hydraulic servo of the band brake during gear shifting. However, in order to provide a smooth engagement force to the other frictional engagement devices mentioned above, it is not possible to change the working oil pressure during gear shifting, and in the past, the pressure receiving surface on the engagement side was single, so the area was the same and the pressure receiving surface was the same. The working oil pressure during gear shifting is determined by adjusting the ratio to the pressure receiving area on the side, and if you simply increase the pressure receiving area on the engagement side, the pressure receiving area on the disengaging side must also increase proportionately. However, the disadvantage is that the hydraulic servo becomes larger and requires major equipment changes.

The present invention can easily increase the supporting torque capacity of the band brake without increasing the radius of the band brake, the hydraulic servo, etc., and without changing the magnitude of the hydraulic pressure acting on the pressure receiving surface on the release side during gear shifting. The hydraulic control device for an automatic transmission of the present invention has been proposed for the purpose of changing and increasing the number of input and output shafts by selectively engaging a plurality of hydraulically operated friction engagement devices. In automatic transmissions in which multiple drive ratios are achieved, a band brake is used as one of the frictional engagement devices, and the hydraulic piston of the hydraulic servo that operates the band brake has two independent hydraulic pistons on the engagement side. Two pressure receiving surfaces are formed on the engagement side and one pressure receiving surface is formed on the release side, and during gear shifting, two pressure receiving surfaces are formed on the engagement side.
Hydraulic pressure acts on one of the two pressure-receiving surfaces, and when a certain gear stage is achieved by engagement of the band brake, oil pressure acts on both sides of the two pressure-receiving surfaces on the engagement side.

An example in which the present invention is applied to a vehicle automatic transmission with three forward speeds and one reverse speed is shown in FIGS. 1 to 6 below.
This will be explained in detail according to the drawings.

In FIG. 1, a crankshaft 10 driven by an engine (not shown) is connected to a pump 16 of a torque converter 14 via a casing 12, and a stator 1 of the torque converter 14.
8 is connected to a transmission casing 22 via a one-way clutch 20. Further, the turbine 24 of the torque converter 14 is connected to a clutch 28 and a clutch 30 via an input shaft 26, and the output side of the clutch 28 is connected to an intermediate shaft 32.
The output side of the clutch 30 is connected to the rear sun gear 36 of a Ravigneau planetary gear unit 34 (hereinafter simply referred to as a gear unit) through a sleeve shaft 38 fitted externally to the intermediate shaft 32, and the output side of the clutch 30 is connected to the front side of the gear unit via a sleeve shaft 38 externally fitted on the intermediate shaft 32. It is connected to a sun gear 40 and also to a brake 42 . The gear device 34 includes the rear sun gear 36, the front sun gear 40, a long pinion gear 46 rotatably supported on a rotatably arranged planetary carrier 44, a short pinion gear 48, and a ring gear 50. The pinion gear 46 is meshed with a front sun gear 40, a short pinion gear 48, and a ring gear 50, the rear sun gear 36 is meshed with a short pinion gear 48, and the ring gear 50 is connected to an output shaft 52. The planetary carrier 44 is also connected to the transmission casing 22 via a one-way clutch 54 and to a brake 56.

The one-way clutch 54 is provided to prevent the planetary carrier 44 from being reversed.

Each of the above-mentioned crunches and brakes is equipped with an engaging piston device or servo device, etc.
Engagement and disengagement are performed by supplying and discharging hydraulic pressure, and hydraulic pressure is selectively supplied and discharged to the piston device, servo device, etc. by actuation of a shift valve etc. installed in the hydraulic circuit, and each clutch and brake is activated. The coefficient of
By combining the release, three forward speeds and one reverse speed are achieved.

The table in Figure 2 shows the relationship between the operation of each of the clutches and brakes mentioned above and the gear status in each range of the manual valve. The "one" mark indicates their release state. As is clear from the table, the clutch 28 is always engaged when achieving a forward drive gear, and the one-way clutch 54 is engaged when the manual valve is in the D or 2 range. The clutch 30 is engaged only when the gear position (1st) is to be achieved, and the clutch 30 is engaged when the high speed gear (3rd) is to be achieved. Further, the brake 42 is engaged when achieving a medium speed gear (2nd), and the clutch 30 and brake 5 are engaged when moving in reverse.
6 is engaged, and the 1st operation when the manual valve is in the L range is achieved by engaging the brake 56 instead of the one-way clutch 54 to enable engine braking.

FIG. 3 shows a supply hydraulic circuit for the friction engagement device particularly related to the present invention among the four hydraulic friction engagement devices in the power train, namely the brake 42 that uses a band brake and the clutch 30 that uses a clutch disk. As shown, clutch 3
0 and the brake 42 are supplied from an oil pump (not shown) through various control valves.

FIG. 4 is an enlarged sectional view of the hydraulic servo 56 of the brake 42 shown in FIG. piston 58
The tip of the protruding portion of the rod 60 fitted in the center of the piston 58 is slidably fitted to the brake drum 6 which rotates integrally with the front sun gear 40.
The brake band 62 is in contact with an engaging portion 66 fixed to an end of a brake band 64 that frictionally engages with the outer circumferential surface of the drum 2 to fix the drum 62 to the casing 22.

A hydraulic chamber 70 surrounded by the release side pressure receiving surface 68 of the hydraulic piston 58 and the casing 22 communicates with an oil passage 74 formed in the transfer plate 72.

Further, a spring 76 is installed in the hydraulic chamber 70 to press the hydraulic piston 58 toward the release side. On the other hand, the engagement side pressure receiving surface of the hydraulic piston 58 is 2
The outer peripheral surface of the annular step and the casing 2
A sleeve 80 is fitted between the inner circumferential surfaces of the two, the sleeve 80 is fixed to the casing 22 by a snap spring 82, and the hydraulic piston 58 slides on the inner circumferential surface of the sleeve 80.

Furthermore, a lid 8 is provided in the hollow cylindrical portion of the casing 22.
4 is fitted, and the lid 84 is fixed to the casing 22 by a snap spring 86.

The outer pressure receiving surface 88 of the two divided engagement side pressure receiving surfaces of the hydraulic piston 58 forms a hydraulic chamber 90 with the sleeve 80, and the hydraulic chamber 90 is connected to an oil passage formed in the transfer plate 72. 92 , and the inner pressure receiving surface 94 is connected to the lid 84 to form the hydraulic chamber 9 .
6, and the hydraulic chamber 96 communicates with an oil passage 98 formed in the transfer plate 72.

The oil passage 98 is communicated with an oil passage 102 via a servo control valve 100, and the oil passage 92 is directly connected to the oil passage 102.

The oil passage 74 connects to the oil passage 10 via the orifice 104.
6, the oil passage 106 is connected to an orifice 108,
It communicates with a hydraulic chamber (not shown) of the clutch 30 via a check valve 110.

The servo control valve 100 has a spool 116 and a spring 118, which have opposing pressure receiving surfaces 112 and 114. When the spool 116 is at the left end position shown in the figure, it connects the oil passage 102, and when it is switched to the right end position, it drains the oil passage 98. It communicates with the oil passage 120.

The oil pressure of the oil passage 92 acts on the pressure receiving surface 112, and the oil pressure of the oil passage 74 acts on the pressure receiving surface 114.

Hydraulic pressure is selectively supplied to the oil passage 102 and the oil passage 106 according to the operating state by various switching valves (not shown), and the oil pressure is adjusted to a constant pressure by a pressure regulating valve (not shown). line pressure.

In this embodiment, as is clear from FIG. 2, oil pressure is supplied to the oil passage 102 when the middle gear is achieved, and oil pressure is supplied to the oil passage 106 when the high gear is achieved. In addition, an orifice, an accumulator, etc. (not shown) are interposed in the hydraulic pressure supply oil passages to the oil passages 102 and 106, and the oil passages 102 and 106 are provided with an orifice, an accumulator, etc.
Even when oil is supplied to the oil pumps 02 and 106, the pressure does not immediately rise to the line pressure, but reaches the line pressure with a delay, so that the clutch 30 and the brake 42 can be engaged and released smoothly without any shock.

Hydraulic piston 5 of brake 42 in FIG.
8 is at the left end position, indicating a state in which the brake 42 is engaged and a middle gear is achieved.

That is, the oil pressure in the oil passage 106 is discharged, and line pressure is supplied to the oil passage 102.

The line pressure supplied to the oil passage 102 is guided from the oil passage 92 to the hydraulic chamber 90 of the brake 42, and is also guided from the oil passage 98 to the hydraulic chamber 96 of the brake 42 through the servo control valve 100.
It also acts on the pressure receiving surface 112 of the servo control valve 100.

The line pressure supplied to the oil passage 106 is guided from the oil passage 74 to the hydraulic chamber 70 of the brake 42 and also acts on the hydraulic chamber of the clutch 30 and the pressure receiving surface 114 of the servo control valve 100, but when the middle gear is achieved, The oil is drained, and the spool 116 of the servo control valve 100 is connected to the pressure receiving surface 1.
The oil passage 102 and the oil passage 98 are in communication with each other because of the pressing force of the line pressure acting on the oil passage 102 and the oil passage 98, which is at the left end position against the biasing force of the spring 118.

The engagement of the brake 42 at this time is such that the pressing force of the hydraulic piston 58 is F and the biasing force of the spring 76 is F ₁
The area of the pressure receiving surface 88 is A ₁ and the area of the pressure receiving surface 94 is
A ₂ , and the line pressure is _PL , F=(A ₁ +A ₂ ) _PL −F ₁ This is achieved with the pressing force expressed by the formula (1).

In order to achieve a high speed stage from this state, the oil passage 10
When line pressure is supplied to 6, the line pressure is guided to the hydraulic chamber 70 of the brake 42, the clutch 30, and the servo control valve 100, but oil is gradually supplied by the action of an orifice (not shown). Due to the accumulator effect of the hydraulic servo 56, the working oil pressure during gear shifting is maintained at a low oil pressure.

This pressure reduction during gear shifting is important for smooth engagement of the clutch 30 and release of the brake 42 with overlap, and for reducing gear shift shock.
2 and the torque fluctuation is explained with reference to FIG. 5. Let P ₂ be the low oil pressure required to achieve the smoothest engagement of the clutch 30, and as shown by the line pressure shown as a solid line P _L in FIG. When a low oil pressure of approximately P ₂ occurs during shifting, smooth fluctuations in the output shaft torque as shown by the solid line T are obtained, but when P _L becomes higher than P ₂ during gear shifting, as shown by the broken line in Fig. 5. If the fluctuation of the output shaft torque becomes large and a shift shock occurs, and if P _L becomes lower than P ₂ during the shift, the shift time becomes longer and the shift feeling deteriorates. The above P ₂ is determined by the balance of forces acting on the hydraulic piston 58 of the hydraulic servo 56.

In this embodiment, when a hydraulic pressure of P ₂ is generated in the oil passage 106, P ₂ also acts on the pressure receiving surface 114 of the servo control valve 100, and the spool 11 is immediately
6 is switched to the right end position in FIG.
is connected to an oil drain path 120.

Therefore, during gear shifting, the balance of forces acting on the hydraulic piston 58 is expressed by the formula A ₃ P ₂ + F ₁ = A ₁ P _L , where A ₃ is the area of the pressure receiving surface 68, and P ₂ = A ₁ /A ₃ ×P−F ₁ /A ₃ ...Equation (2) is obtained.

As is clear from the above, in the past, the hydraulic piston 58 was engaged and operated by the line pressure P _L acting on only the pressure receiving surface 88, but in this embodiment, the hydraulic piston 58 is released. During gear shifting, the line pressure P _L acting on the pressure receiving surface 94 is discharged under the control of the servo control valve 100 to obtain the same line pressure P ₂ during gear shifting as in the past, and smooth gear shifting is achieved. When the brake 42 is engaged, by applying line pressure to both pressure receiving surfaces 88 and 94, the pressing force F of the hydraulic piston 58 is increased, and as a result, the torque capacity of the brake 42 is increased compared to the conventional one. In addition, the above formula (1) and
As is clear from equation (2), in this embodiment, the hydraulic piston 58 can be moved by simply changing the area A ₂ of the pressure receiving surface 94 without changing the line pressure P ₂ during shifting, which is important for reducing the shifting shock. It is possible to change the pressing force F.

Next, referring to FIG. 6, a transmission means having an area of A ₂ will be explained. The hollow cylindrical part formed in the transmission casing 22 has a two-stage cross section with two inner peripheral surfaces 122 and 124 having different radii on the circumferential surface. If the annular sleeve 126 of the sleeve 126 is fitted and the hydraulic piston 58 slides in contact with the inner circumferential surfaces 122, 124, and the radius of the inner circumferential surfaces 122, 124 of the sleeve 126 is appropriately set, P ₂ , the area of the pressure receiving surface 94 can be reduced to an arbitrary size.

Therefore, according to the present embodiment, there is no need to change the line pressure P _L that would cause problems such as deterioration of the reliability of the hydraulic control mechanism and reduction in hydraulic transmission efficiency, and the band change that would require major equipment changes can be avoided. The shapes of the hydraulic piston 58 and the sleeve 80 can be changed without changing the band radius of the brake 42 or the radius of the hydraulic servo 56, and without changing the line pressure _P2 during shifting, which is important for reducing the shifting shock. This has the effect that a band brake having a supporting torque capacity suitable for various engines with different engine outputs can be obtained with only slight changes.

Incidentally, in the above embodiment, the supply and discharge control of the line pressure P _L acting on the pressure receiving surface 94 is performed using a servo control valve operated by the oil pressure acting on the pressure receiving surface 68 of the hydraulic piston 58 and the oil pressure acting on the pressure receiving surface 88. As shown in FIG. 7, this control is performed by detecting the movement of the hydraulic piston 58 of the hydraulic servo 56 by a switch 128, and interlocking the opening/closing operation of the switch 128 with a hydraulic control solenoid valve 130. valve 130
It is also possible to use a servo control valve 100 that switches according to the operation of the servo control valve 100.

In FIG. 7, parts that are substantially the same or equivalent to those of the embodiment shown in FIG. 3 are given the same reference numerals. In this modification, a pressure receiving surface 112 is formed only on one end surface of the spool 116 of the servo control valve 100, and when hydraulic pressure acts on the pressure receiving surface 112, the spool 116 resists the urging force of the spring 118 and moves as shown in FIG. As shown, the oil passage 102 and the oil passage 98 are connected to each other, and the pressure receiving surface 1 is moved to the left end position as shown.
When the hydraulic pressure of the oil pump 12 decreases or becomes zero, the spool 116 moves to the right end due to the biasing force of the spring 118, and the oil passage 98 communicates with the oil drain passage 120.

Hydraulic pressure is supplied to the pressure receiving surface 112 via an orifice 134 from an oil passage 132 to which line pressure is supplied, and the orifice 134 and the pressure receiving surface 112
When the orifice 136 provided between the
When the line pressure P _L acts on the pressure receiving surface 112 and the orifice 136 is open, the oil pressure acting on the pressure receiving surface 112 from which oil is drained becomes zero.

Solenoid valve 130 is connected in series to switch 128 and battery 138, and
8 is configured to open only when the hydraulic piston 58 moves to the right.

[Brief explanation of the drawing]

Fig. 1 is a gear train diagram showing one embodiment of the present invention. Fig. 2 is a diagram showing the relationship between the operation of each friction engagement device and the gear position status in the same embodiment. Fig. 3 is a hydraulic circuit diagram of the same embodiment. , FIG. 4 is an enlarged sectional view of the hydraulic servo adopted in the same embodiment, FIG. 5 is a diagram showing the relationship between line pressure P _L and output shaft torque T during gear shifting in the same embodiment, and FIG. FIG. 7 is an enlarged sectional view of a hydraulic servo showing a modification of the same embodiment, and FIG. 7 is a hydraulic control diagram showing a modification of the present invention. 30...clutch, 42...brake, 56...
...Hydraulic servo, 58...Hydraulic piston, 68,8
8, 94...pressure receiving surface, 100...servo control valve.

Claims (1)

[Claims] 1. In an automatic transmission in which a plurality of drive ratios are achieved between input and output shafts by selectively engaging a plurality of hydraulically operated friction engagement devices, one of the friction engagement devices A band brake is used, and the hydraulic piston of the hydraulic servo that operates the band brake has two independent pressure receiving surfaces on the engagement side and one on the release side.
During gear shifting, hydraulic pressure acts on one of the two pressure receiving surfaces on the engaging side, and when a certain gear is achieved by engaging the band brake, the two pressure receiving surfaces on the engaging side A hydraulic control device for an automatic transmission characterized by hydraulic pressure acting on both sides. 2 Switch control of the supply and discharge of the hydraulic pressure acting on the other pressure receiving surface on the engagement side using a servo control valve operated by the hydraulic pressure acting on the pressure receiving surface on the release side and the hydraulic pressure acting on one pressure receiving surface on the engagement side. A hydraulic control device for an automatic transmission according to claim 1. 3. A patent that detects the movement of a hydraulic piston toward the engagement side or the disengagement side and operates a servo control valve in response to the movement to switch and control the supply and discharge of hydraulic pressure acting on one pressure receiving surface on the engagement side. A hydraulic control device for an automatic transmission according to claim 1.