JPH07180752A - Automatic transmission for vehicle - Google Patents

Abstract

PURPOSE:To actuate gear shifting for four forward speeds only by two epicyclic gear mechanisms without switching an input shaft, and moreover reduce the number of one-way clutches. CONSTITUTION:The ring gear R111 of a single pinion type epicyclic gear mechanism 111 is connected to a sun gear S111 by a second clutch K112 via a first one-way clutch F111, and the ring gear R111 is fixed by a first brake B111 via the same one-way clutch F111. Besides, the sun gear S111 is also fixed by the same first brake B111 via a second clutch K112. Meanwhile, the sun gear S121 of a double pinion type epicyclic gear mechanism 121 is fixed by a second brake 112. By this constitution, when gear shifting is actuated from first forward speed to second forward speed and from third forward speed to fourth forward speed, the first one-way clutch is double used, and the one-way clutch can thereby be reduced by one in number.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an automatic transmission for vehicles, and more particularly to an automatic transmission for vehicles having two sets of planetary gear mechanisms.

[0002]

2. Description of the Related Art As a stepped automatic transmission for a vehicle, many automatic transmissions having a gear train mainly composed of a plurality of sets of planetary gear mechanisms have been proposed.

In this type of automatic transmission, if the number of planetary gear mechanisms to be used is increased, the number of shift stages that can be set increases, but on the other hand, the automatic transmission becomes large or heavy. It causes the problem of increase.

It is preferable that the gear ratio that can be set has a relationship close to a geometric series, but in order to do so, the gear ratio (the ratio of the number of teeth of the sun gear to the number of teeth of the ring gear) is normal. If a larger planetary gear mechanism must be used, productivity, ease of assembly, durability, etc. may be impaired.

As described above, there are many contradictory elements in the automatic transmission for a vehicle, and it is very difficult to have a structure that satisfies all the elements. Among conventional demands for automatic transmissions, there is an example in which two sets of planetary gear mechanisms are used as a gear train for an automatic transmission that has a large number of gear stages that can be set and that can be downsized. JP-A-2-2299
No. 46 publication.

The structure will be briefly described with reference to FIG. 10. The first planetary gear mechanism PG1 includes a sun gear S1, a ring gear R1, a short pinion PS1 meshing with the gears S1 and R1, and the short pinion. PS1
The second planetary gear mechanism PG2 includes a sun gear S2, a ring gear R2, and a short pinion PS2 that meshes with the ring gear R2, and also has a long pinion PL. PL meshes with the sun gear S2 and the short pinion PS2. The pinions PS1, PL, PS2 are held by a common carrier PC.

On the other hand, as friction engagement devices, first to third clutches K1, K2, K3 and first to third brakes B are provided.
1, B2, B3 are provided. The first clutch K1 is provided between the input shaft IS and the second sun gear S2, the second clutch K2 is provided between the input shaft IS and the first ring gear R1, and the third clutch K3 is the input shaft IS. And the first sun gear S1.

On the other hand, the first brake B1 is provided between the second sun gear S2 and the housing HU, and
The brake B2 is provided between the second ring gear R2 and the housing HU, and the third brake B3 is provided between the first ring gear R1 and the housing HU. or,
The output member OM is connected to the common carrier PC.

In the gear train shown in FIG. 10, each friction engagement device is engaged as shown in FIG.
2 speeds can be set, and 2 speeds can be set in reverse.
It is said that it can be miniaturized because it can be combined. Note that FIG.
The circles indicate engagement and the blanks indicate release.

[0010]

SUMMARY OF THE INVENTION In the above-described conventional automatic transmission provided with a gear train, the fourth speed, which is a direct connection stage having a gear ratio of "1", is sandwiched between the fourth speed and the lower gears. Since the third clutch K3 is engaged and the second clutch K2 is engaged at the high speed side gear stage, when the automatic transmission is configured as an overdrive stage having a gear ratio smaller than "1", Shifting between the overdrive stage and the underdrive stage causes so-called input clutch switching. Therefore, in the above-described conventional automatic transmission equipped with a gear train, there is a possibility that a shift shock accompanying switching of the input clutch may occur and the riding comfort may be impaired, and the shift shock accompanying switching of the input clutch is suppressed. Therefore, a high degree of shift control is required, and the cost of the control device may increase.

Against the background of such circumstances, the applicant previously disclosed in Japanese Patent Application No. 4-283659 (unknown) that a gear train mainly composed of two sets of planetary gear mechanisms is used to switch the input clutch. We proposed an automatic transmission for vehicles that can set four forward speeds including overdrive speeds without performing it.

This automatic transmission has a sun gear S12, S22, a ring gear R12, and a sun gear S12, as shown in FIG.
R22 and a carrier C2 that holds the pinion gears P12 and P22 are provided with a gear train mainly composed of two sets of first and second planetary gear mechanisms 12 and 22 having rotating elements.
Of these first and second planetary gear mechanisms 12 and 22, (at least one) pinion gear P12 in (one) the first planetary gear mechanism 12 and (at least one) in (the other) second planetary gear mechanism 22. Pinion gear P22
Are integrally formed with each other to form a long pinion, and the carriers C2 of the planetary gear mechanisms 12 and 22 are commonly rotated together with the long pinion.

Therefore, since the carrier C2 of the two sets of the first and second planetary gear mechanisms 12, 22 each having three rotating elements is integrated, the gear train has five rotating elements. Will be.

FIG. 14 is a nomographic chart showing the shift characteristics of the automatic transmission shown in FIG. 12, and FIG. 13 is an example of an operation table. The alignment chart is created on the same basis as the alignment chart shown in the above-mentioned publication. That is, in this automatic transmission, since the carrier C2 is shared as described above, there are five rotating elements, and therefore the alignment chart is shown by five vertical lines. In addition, the first planetary gear mechanism 12
Is a single pinion type planetary gear mechanism, the second planetary gear mechanism 22 is a double pinion type, and the pinion gear P22 on the outer peripheral side of the second planetary gear mechanism 22 is a long pinion. On one side of the line indicating C2, the sun gear S of the first planetary gear mechanism 12
12 is located, and a line showing the ring gear R22 of the second planetary gear mechanism 22, a line showing the sun gear S22, and a line showing the ring gear R12 of the first planetary gear mechanism 12 are located on the other side. Become. Each planetary gear mechanism 1
The line showing the ring gear R22 of the second planetary gear mechanism 22 is located adjacent to the line showing the carrier C2 in the relationship of the gear ratios of 2 and 22, and the output member (32 or the ring gear R2
2). In addition, the ring gear R of the first planetary gear mechanism 12
The line indicating 12 is located adjacent to the line indicating the ring gear R22 which is the output member. Then, a straight line connecting the position of the length "1" on the line of the ring gear R22 of the second planetary gear mechanism 22 which is the output element and the origin position of the line showing the fixed element is the input member (42, or In the forward stage, the position crossing the line indicating the carrier C2) is the input rotation speed, so
The value of the input rotation speed becomes the gear ratio.

In FIG. 12, reference numerals K02 and K1.
2, K22 is a clutch, B02, B12, B22, B3
2 is a brake, F02, F12 and F22 are one-way clutches, 31 is an output shaft, 42 is an input shaft, and 52 is a casing.

Further, FIG. 13 is an operation table, and the circle marks indicate engagement,
The blank column indicates release, the Δ symbol indicates engagement in a state not involved in power transmission, and the ⊚ symbol indicates engagement during engine braking.

By the way, a new problem has arisen when the automatic transmission according to this unknown prior invention is attempted to be specifically set.

That is, in this automatic transmission, the first
In addition, since one one-way clutch is used for each shift, there is a problem in that a space is required in the axial direction or the radial direction, the weight is increased, and the cost is also increased. Since the difference in rotational speed between the adjacent sun gear members S12 and S22 is large, the durability of the thrust bearing arranged so as to be sandwiched between the two members S12 and S22 is likely to be impaired with respect to other bearings. was there.

The present invention has been made in view of the above circumstances, and the number of one-way clutches can be reduced by using one one-way clutch redundantly at two shifts. By providing an automatic transmission for a vehicle that can reduce the accommodation space, reduce the weight, and reduce the cost, and can eliminate the problem caused by the large difference in the relative rotational speeds of both sun gear members, It is a solution to the above problems.

[0020]

According to a first aspect of the present invention, a sun gear, a ring gear arranged concentrically with respect to the sun gear, and at least one pinion gear arranged between these gears, A carrier that holds the pinion gear, and two sets of planetary gear mechanisms that have rotating elements are provided. The carriers are connected so as to rotate integrally with each other, and among the two sets of planetary gear mechanisms, At least one pinion gear in the single pinion type planetary gear mechanism and at least 1 in the double pinion type planetary gear mechanism, wherein one planetary gear mechanism is constituted by a single pinion type planetary gear mechanism and the other planetary gear mechanism is constituted by a double pinion type planetary gear mechanism. For vehicle use in which two ring gears and a pinion gear that meshes are integrated to form a long pinion. In a transmission, a pinion that meshes with a sun gear of a double pinion type planetary gear mechanism is connected to the carrier, a link gear of the double pinion type planetary gear mechanism is connected to an output shaft, and a connection state of each rotating member is selected. As a control element, a first clutch that always connects the carrier and the input shaft in the forward stage, and a first clutch that connects the ring gear and the sun gear of the single pinion type planetary gear mechanism through the first one-way clutch. Two
A clutch and a ring gear of the single-pinion type planetary gear mechanism are fixed to the casing via the first one-way clutch, and a sun gear of the single-pinion type planetary gear mechanism is fixed to the casing via the second clutch. The above problem is solved by providing one brake and a second brake for fixing the sun gear of the double pinion type planetary gear mechanism to the casing.

Further, the invention according to claim 2 solves the above problem by disposing the carrier between the both sun gears, and at the same time, the difference in the rotational speeds of both the sun gears which are also adjacent members. It is intended to solve the problem caused by the large.

According to the third aspect of the present invention, the piston constituting the first brake has a double structure to form a hydraulic chamber in each, and hydraulic oil can be independently supplied to and discharged from each hydraulic chamber. With the configuration, the above problem is also solved, and in addition to using the first one-way clutch twice, the brake that has to be used twice at different speeds (first
In view of the fact that the (brake) comes out, a problem that newly occurs when the first brake is used for shifting twice is also eliminated.

[0023]

In the present invention (claims 1 to 3), all forward gears can be achieved without switching the input clutch, as in the previously unknown invention.

The shift from the first speed to the second speed is achieved by newly engaging only the second brake. In this case, the first one-way clutch arranged between the ring gear and the sun gear of the single pinion type planetary gear mechanism changes from the engaged state to the idling state, and so-called one-way clutch-to-brake smooth shifting is realized. .

On the other hand, in the present invention, the shift from the third speed to the fourth speed is basically realized by newly engaging only the first brake. At this time, the same first one-way clutch is also changed from the engaged state to the released state, and a smooth one-way clutch-to-brake shift is realized here as well.

As a result, the one-way clutch, which was provided for each shift in the configuration according to the previously unknown invention, can be surely reduced by one while having the same function, and the corresponding amount can be accommodated. Space can be reduced, weight can be reduced, and cost can be reduced.

The structure of the automatic transmission according to the present invention is as follows.
There is a tendency that the difference in rotation speed between the two sun gear members, which are members adjacent to each other, becomes large. Therefore, the durability of the thrust bearing arranged so as to be sandwiched between both members is likely to be impaired with respect to other bearings. In the invention described in claim 2, since the carrier that is the input member during forward movement is arranged between both sun gears, the relative rotational speed of both sun gear members can be reduced, and the durability of the bearing can be improved. .

Further, in the automatic transmission according to the present invention,
Since the one-way clutch is used twice, the first brake is the first
It will be used repeatedly at the 4th speed and the 4th speed. However, the torque capacity required for the first brake is different between the first speed and the fourth speed. Therefore, if the first brakes are used in duplicate without any allowance, it is necessary to increase the pressure receiving area so as not to cause trouble even at the first speed, which requires a larger torque capacity. In particular, in the shift from the third speed to the fourth speed, there is a problem that the capacity becomes too large and the shift shock becomes large. In order to avoid this problem, it is necessary to perform a separate control for reducing the engagement force during the shift from the third speed to the fourth speed, but the piston with a large pressure receiving area is finely controlled with a small hydraulic pressure. Is very difficult in reality.

Further, in the automatic transmission according to the present invention,
When shifting from the first speed to the third speed and from the fourth speed to the third speed, it is necessary to quickly shift the first brake from the engaged state to the disengaged state, but if the hydraulic chamber is large. The piston returns slowly, resulting in a long shift time and a large shock.

According to the third aspect of the present invention, since the piston constituting the first brake has a double structure and a hydraulic chamber is formed in each of them, the hydraulic oil can be independently supplied to and discharged from the hydraulic chamber. The hydraulic oil is supplied to the two hydraulic chambers at the same time when the first speed is engaged, which requires a large capacity, and the hydraulic oil is supplied to and discharged from only one hydraulic chamber, when the fourth speed is engaged, which does not require a large capacity. As a result, this problem can be reasonably solved.

[0031]

Embodiments of the present invention will now be described in detail with reference to the drawings.

1 to 3 show a first embodiment of the present invention.

In this embodiment, the first planetary gear mechanism 111 is constituted by a single pinion type planetary gear mechanism, and the second planetary gear mechanism 121 is constituted by a double pinion type planetary gear mechanism.

The first planetary gear mechanism 111 includes a sun gear S1.
11, a ring gear R111, and a carrier C101 that holds a pinion gear P111 meshed with these are used as rotating elements.

On the other hand, in the second planetary gear mechanism 121, between the sun gear S121 and the ring gear R121, the short pinion gear PS121 that meshes with the sun gear S121.
And a pinion gear P121 that meshes with the short pinion gear PS121 and the ring gear R121. The pinion gear P121 is integrated with the pinion gear P111 of the first planetary gear mechanism 111 to form a long pinion. Further, the carrier C101 for the first planetary gear mechanism 111 and the second planetary gear mechanism 121 is shared. The drive gear 131, which is an output member, is integrally connected to the ring gear R121 of the second planetary gear mechanism 121.

The input shaft 141 has a carrier C101 and a second
It is selectively connected to the sun gear S121 of the planetary gear mechanism 121. As the friction engagement device therefor, the input shaft 141
And a carrier C101, a first clutch K111 is provided.

Further, the sun gear S of the first planetary gear mechanism 111.
A second clutch K112 and a first one-way clutch F111, which are arranged in series with each other, are arranged between 111 and the ring gear R111.

Further, a third clutch K113 is provided between the input shaft 141 and the sun gear S121 to secure the engine brake at the third speed and the reverse range.

On the other hand, as a braking means, the first brake B is connected in series with the second clutch K112 between the sun gear S111 of the first planetary gear mechanism 111 and the housing 151.
111 is provided. The member on the clutch side of the first brake B111 is the first member via the one-way clutch F111.
It is connected to the ring gear R111 of the planetary gear mechanism 111. This ring gear R111 is the third brake B113.
It is connected to the casing 151 via.

Further, the sun gear S1 of the second planetary gear mechanism 121.
21 and the housing 151 between the second brake B11
Two are provided.

A fourth brake B114 and a second one-way clutch F112, which are arranged in series with each other, are arranged in parallel with the second brake B112.

Incidentally, this second one-way clutch F112.
Are set to engage when the sun gear S121 tries to rotate in the reverse direction.

FIG. 3 is a collinear diagram showing the shift characteristics of the automatic transmission shown in FIG. The operation table is shown in FIG.

In this automatic transmission, the carrier C1
Since 01 is common, there are five rotating elements, so the alignment chart is shown by five vertical lines. Also, the first planetary gear mechanism 111 is a single pinion type planetary gear mechanism, the second planetary gear mechanism 121 is a double pinion type, and the pinion gear P121 on the outer peripheral side of the second planetary gear mechanism 121 is a long pinion. Therefore, on one side of the line showing the common carrier C101,
The line showing the sun gear S111 of the planetary gear mechanism 111 is located, and the line showing the ring gear R121 of the second planetary gear mechanism 121 and the line showing the sun gear S121 are located on the other side.
The line indicating the ring gear R111 of the planetary gear mechanism 111 is positioned.

The ring gear R1 of the second planetary gear mechanism 121 is related to the gear ratio of the planetary gear mechanisms 111 and 121.
The line indicating 21 is located adjacent to the line indicating the carrier C101 and serves as an output member. Further, the first planetary gear mechanism 11
The line indicating the first ring gear R111 is located adjacent to the line indicating the ring gear R121 that is the output member. A straight line connecting the position of length "1" on the line of the carrier C101 which is the input element at the time of forward movement and the origin position of the line showing the fixed element is the output member (ring gear R121 of the second planetary gear mechanism 121). Since the position crossing the line indicating is the output rotation speed, the value of this rotation speed is the output rotation speed per input rotation.

Here, each shift speed set in the automatic transmission according to the first embodiment will be described.

First, the first speed is set by engaging the first one-way clutch F111 with the engagement of the first clutch K111 and the first brake B111. That is, when the carrier 101 rotates together with the input shaft 141, the load on the output side is applied to the ring gear R121 of the second planetary gear mechanism 121, so the first planetary gear mechanism 11
The first ring gear R111 tries to rotate in the reverse direction, but when torque in the reverse rotation direction is applied to the ring gear R111, the first one-way clutch F111 is engaged. As a result, the rotation of the ring gear R111 is blocked (that is, the ring gear R11
(1 is given a reaction torque), and the ring gear R121 of the second planetary gear mechanism 121, which is an output member, is greatly decelerated with respect to the input shaft 141 and rotates normally.

Regarding the gear ratio at the first speed, the gear ratio of the first planetary gear mechanism 111 is ρ12, the gear ratio of the second planetary gear mechanism 121 is ρ22, and the pinion gear and ring gear on the outer peripheral side of the second planetary gear mechanism 121 are the same. When the tooth number ratio is γ22, it is represented by (1-ρ12) / (1-ρ12-2 · γ22) (1).

In order to apply the engine brake at the first speed, the third brake B113 provided in parallel with the first one-way clutch F111 is engaged, so that the first one-way clutch F111 runs in the idling direction. Ring gear R111
Prevent the rotation of.

On the other hand, in the second speed, the second clutch is engaged with the engagement of the first clutch K111 and the fourth brake B114.
It is set by engaging the one-way clutch F112 and idling the first one-way clutch F111. That is, when the carrier C101 rotates together with the input shaft 141, a load on the output side is applied to the ring gear R121 of the second planetary gear mechanism 121, and the sun gear S121 tries to rotate in the reverse direction. Therefore, the second one-way clutch F112 is engaged and the sun gear S121 is engaged. S121 is fixed (reaction torque is applied to the sun gear S121), and as a result, the ring gear R121, which is an output member, rotates the input shaft 141 by rotating the second planetary gear mechanism 1.
It rotates forward at the rotation speed connected at 21. Therefore, this second
The speed change ratio is represented by 1 / (1-ρ22) (2). In this case, the first one-way clutch F1
11 changes from the engaged state up to that point to the idling state, and allows the normal rotation of the ring gear R111 of the first planetary gear mechanism 111.

This second speed is the second one-way clutch F112.
In order to prevent the rotation of the sun gear S121 in the idling direction of the second one-way clutch F112 when the engine brake is applied,
The second brake B112 is engaged.

If the so-called clutch-to-clutch shift is allowed, the fourth brake B114 and the second one-way clutch F112 are not always necessary.
The skeleton with these omitted will be described later with reference to FIGS. 4 and 5. Since the fourth brake B114 and the second one-way clutch F112 are provided in the embodiments of FIGS. 1 to 3, when shifting from the first speed to the second speed, the fourth brake B114 is newly engaged. It is possible to achieve a smooth shift by means of a one-way clutch and two brakes with less shift shock.

In the third speed, the first clutch K111 and the second clutch
It is set by engaging the first one-way clutch F111 again as the clutch K112 is engaged. That is, by engaging the second clutch K112, the first clutch
The sun gear S111 of the planetary gear mechanism 111 is rotated forward and the first one-way clutch F111 is engaged. As a result, in the first planetary gear mechanism 111, the carrier C101, the sun gear S111, and the ring gear R111 rotate together with the input shaft 141. The whole rotates together. Therefore, the ring gear R121, which is an output member, rotates forward at the same speed as the input shaft 141, and is in a direct connection stage having a gear ratio of "1".

In this case also, the first one-way clutch F
Since 111 is engaged, the third clutch K113 is engaged when the engine brake is applied. or,
The second one-way clutch F112 is the second
Until the sun gear S121 of the planetary gear mechanism 121 normally rotates at the third speed, the idling state is maintained. Therefore, even if the fourth brake B114 is engaged at the third speed, the fourth brake B11 is engaged.
4 is not involved in power transmission. Therefore, the fourth brake B
Since 114 does not require synchronous release for shifting, shifting from the second speed to the third speed requires the second clutch K112.
Can be achieved only by newly engaging.

The fourth speed is set by engaging the first clutch K111, the second clutch K112, and the first brake B111. That is, since the carrier C101 is commonly used by the first planetary gear mechanism 111 and the second planetary gear mechanism 121, the carrier C101 is the same as the carrier C101.
When rotated together with the first planetary gear mechanism 111,
The sun gear S111 fixed by the brake B111 and the second clutch K112 receives a reaction torque, and as a result, the common carrier C101 rotates forward around the sun gear S121, and the ring gear R121 of the second planetary gear mechanism 121 is rotated. It is added to the input shaft 141 to rotate forward. That is, an overdrive stage having a gear ratio smaller than "1" is achieved. The gear ratio in this case is represented by (1-ρ12) / (1-ρ12 + 2 · ρ12 · γ22) (3).

Further, in this case, since the sun gear S121 of the second planetary gear mechanism 121 rotates forward faster than the input shaft 141, the one-way clutches F111 and F112 idle even if the fourth brake B114 is engaged. Therefore, the fourth brake B114 does not participate in power transmission. Therefore, the third
The speed change from the fourth speed to the fourth speed can be achieved only by newly engaging the first brake B111, and a smooth speed change is possible.

On the other hand, the reverse speed is set by engaging the third clutch K113 and the third brake B113. That is, the sun gear S121 of the second planetary gear mechanism 121.
Is rotated together with the input shaft 141, the load on the output side is applied to the common carrier C101, and the pinion gear P111 of the first planetary gear mechanism 111 is formed into a long pinion. Although the ring gear R111 of 111 is about to rotate in the forward direction, it is fixed by the brake B111 and receives a reaction torque. As a result, the ring gear R121 of the second planetary gear mechanism 121 rotates in the reverse direction. The gear ratio in this case is represented by (2 · γ22−ρ22 + ρ12 · ρ22) / (ρ22−ρ12 · ρ22-2 · ρ22 · γ22) (4).

As described above, in this automatic transmission, all forward gears can be set while the first clutch K111 remains engaged. Therefore, it is necessary to switch a so-called input clutch (clutch equivalent to the first clutch K111) between the forward gears, as in the conventional example disclosed in Japanese Patent Laid-Open No. 2-229946. Absent. Therefore, increase in shift shock can be prevented, and shift control is facilitated.

Further, as is apparent from the above description, in the automatic transmission according to this embodiment, the first one-way clutch F111 is used for the first speed, the second speed, the third speed, and the fourth speed. Engagement and disengagement are switched between steps. That is, the first one-way clutch F111 is functioning twice (duplicated). As a result, compared to the configuration in which a one-way clutch is arranged for each shift,
At least one one-way clutch can be surely reduced while maintaining the same function, and the storage space, weight, and cost can be reduced accordingly.

FIG. 4 shows a modification of the first embodiment described above.

In this modification, the fourth brake B114 and the second one-way clutch F112 are added to the structure of the first embodiment.
Is further removed. The engagement table of this automatic transmission is as shown in FIG. As is apparent from FIG. 5, in this automatic transmission, the second brake B112 is released and the second clutch K112 is engaged in shifting from the second speed to the third speed. As a result, the second
When shifting from the third speed to the third speed, a so-called clutch-to-clutch shift is performed, but since the fourth brake B114 and the second one-way clutch F112 can be omitted, further reduction of the storage space, weight reduction, and cost reduction are intended. You can

Next, a second embodiment of the present invention will be described with reference to FIGS.

In this second embodiment, the substantially connected state of each rotating member is exactly the same as that shown in FIG. 1, but the arrangement of the carrier C201 is devised so that both sun gears having a large difference in rotational speed. This prevents S211 and S221 from being directly adjacent to each other, thereby preventing the durability of the thrust bearing arranged in this portion from being deteriorated.

FIG. 6 shows the skeleton, and FIG. 7 shows a concrete cross section.

As is clear from both figures, the carrier C20
1 is arranged between the sun gear S211 of the first planetary gear mechanism 211 and the sun gear S221 of the second planetary gear mechanism 221. Therefore, the relative rotation difference between the sliding portions is halved, and it suffices to interpose an inexpensive thrust bearing.

Other configurations are shown in FIGS.
Since it is basically the same as that of the first embodiment shown in FIG. 1, the last two digits in the figure are given the same reference numerals, and duplicate explanations are omitted.

Next, FIG. 8 shows the first brake B21 of FIG.
1 is an enlarged view of the vicinity of 1 (corresponding to the first brake B111 in the case of FIG. 1).

As is apparent from the figure, the first brake B2
11, the piston 211a has a double structure and has oil chambers 211b and 211c, respectively. Each oil chamber 2
The hydraulic oil can be supplied to and discharged from 11b and 211c independently.

The first brake B211 having such a structure allows the oil chamber 21 to be engaged when the first brake B211 is engaged during the shift from the first speed to the second speed.
Hydraulic oil is simultaneously supplied to both 1b and 211c. As a result, the piston 211a can press the friction plate 211d with a large torque with a small time lag, and a reliable fixing action can be obtained.

On the other hand, for example, when engaged during the shift from the third speed to the fourth speed, the hydraulic pressure is supplied to only one of the oil chambers 211b or 211c. As a result, even if the speed itself for supplying the working oil to the oil chamber is not changed at all, the piston 212a uses only the torque corresponding to the decrease in the pressure receiving area as compared with the engagement at the first speed. 211d
Will be pressed. As a result, the friction plate 211d can be pressed with a torque that is necessary and sufficient for shifting from the third speed to the fourth speed.

In the structure of the first brake B211, the piston 211a has a simple shape and does not need to have a large axial dimension, and the number of parts can be increased by adding only one sealing member 211e, so that the cost is increased. With almost no. Further, there is no need to increase the space for the construction of the oil passage, and it is sufficient to supply the pressure to only one oil chamber at the 4th speed, so that the efficiency and the fuel consumption can be improved.

With this configuration, the first one-way clutch F211 (F111) has to be used twice by using the first one-way clutch F211 (F111) twice.
The problem regarding the capacity of the 211 (B111) can be rationally solved.

In the gear train in which the carrier C201 is arranged between the sun gears S211 and S221, the fourth brake B214 and the second one-way clutch F212 are shown in FIG. 9 just as in the previous embodiment. As such, this can be omitted. Since this function and effect have already been described, only the figure is shown and duplicate description is omitted.

[0074]

As described above, according to the present invention,
By using only two sets of planetary gear mechanisms to achieve four forward speeds, it is not necessary to switch the input shaft in each speed change of the forward speed, and by using the one-way clutch for two speed changes, It is possible to omit at least one of them, and it is possible to obtain an excellent effect that the accommodation space, the weight, and the cost can be reduced accordingly.

[Brief description of drawings]

FIG. 1 is a skeleton diagram showing a first embodiment of a vehicle automatic transmission according to the invention.

FIG. 2 is a diagram showing an engagement state of each friction engagement element of the automatic transmission of FIG.

FIG. 3 is a collinear diagram of each rotating member of the above automatic transmission.

FIG. 4 is a skeleton diagram showing a modified example of the first embodiment.

FIG. 5 is a diagram showing an engagement state of each friction engagement element of the modified example.

FIG. 6 is a skeleton diagram showing a second embodiment of the automatic transmission according to the invention.

FIG. 7 is a sectional view showing a specific configuration of the automatic transmission according to the second embodiment.

FIG. 8 is an enlarged sectional view showing the vicinity of the brake B211 in FIG. 7 in an enlarged manner.

FIG. 9 is a skeleton diagram showing a modified example of the automatic transmission according to the second embodiment.

FIG. 10 is a skeleton diagram showing a configuration example of a conventional automatic transmission.

FIG. 11 is a diagram showing an engagement state of each friction engagement element of the conventional example.

FIG. 12 is a skeleton diagram showing a configuration of an automatic transmission corresponding to an unknown prior art.

FIG. 13 is a diagram showing an engagement state of each friction engagement element of the automatic transmission.

FIG. 14 is an alignment chart of each rotating member of the automatic transmission.

[Explanation of symbols]

 111, 211 ... First planetary gear mechanism 121, 221 ... Second planetary gear mechanism S111, S121, S211, S221 ... Sun gear R111, R121, R211, R221 ... Ring gear C101, C201 ... Carrier P111, P211 ... Pinion gear PS121, PS221 ... Short pinion gears P121, P221 ... Pinion gears 141, 241 ... Input shafts 131, 231 ... Output members K111, K211 ... First clutches K112, K212 ... Second clutches F111, F211 ... First one-way clutches B111, B211 ... First brakes B112 , B212 ... second brake 211a ... pistons 211b, 211c ... oil chamber

Claims (3)

[Claims] 1. A rotary element comprising a sun gear, a ring gear arranged concentrically with respect to the sun gear, at least one pinion gear arranged between these gears, and a carrier holding the pinion gear. Two sets of planetary gear mechanisms are provided, the carriers are connected to each other so as to rotate together, and one planetary gear mechanism of the two sets of planetary gear mechanisms is a single pinion type planetary gear mechanism, The other planetary gear mechanism is configured by a double-pinion type planetary gear mechanism, and at least one pinion gear in the single-pinion type planetary gear mechanism and at least one ring gear in the double-pinion type planetary gear mechanism that mesh with each other are integrated and long. In a vehicle automatic transmission that forms a pinion, a double pinion type planetary tooth A pinion that meshes with a sun gear of a car mechanism is connected to the carrier, a link gear of the double pinion type planetary gear mechanism is connected to an output shaft, and the carrier is used as an element for selecting and controlling the connection state of each rotating member. A first clutch that always connects the input shaft with the input shaft in a forward stage; a second clutch that connects the ring gear and the sun gear of the single pinion type planetary gear mechanism through a first one-way clutch; and the single pinion type A first brake for fixing a ring gear of the planetary gear mechanism to the casing via the first one-way clutch, and a sun gear of the single pinion type planetary gear mechanism to the casing via the second clutch; and the double pinion. A second brake for fixing the sun gear of the planetary planetary gear mechanism to the casing, That an automatic transmission for a vehicle. 2. The automatic transmission for a vehicle according to claim 1, wherein the carrier is arranged between the sun gears. 3. The piston according to claim 1 or 2, wherein the piston constituting the first brake has a double structure to form a hydraulic chamber in each, and hydraulic oil can be independently supplied to and discharged from each hydraulic chamber. An automatic transmission for a vehicle, which has a simple structure.