JPH01105042A - Planetary gear type transmission - Google Patents

Abstract

PURPOSE:To obtain a speed change shift of forward five or sixspeed and reverse one-speed by combining a single pinion type planetary gear mechanism with a double-pinion type planetary gear mechanism respectively through three clutches and brakes. CONSTITUTION:A single pinion type planetary gear mechanism 10 and a double- pinion type planetary gear mechanism 20 connect sun gears 11 and 21 by a connecting member 26 and a ring gear 13 to a carrier 25 by a connecting member 27. And by selectively engaging clutches C1-C3 and brakes B1-B3, a speed change shift of forward 5-speed and reverse 1-speed is formed between an input shaft IS and an output shaft OS. That is, the speed change shift obtains the first speed by turning on the clutch C1 and the brake B1, second speed by C2, B2, third speed by C2, B3, fourth speed by C2, C3 and the fifth speed by C3, B3 while the reverse shift by C1, B2. Thus obtaining the forward 5-speed or 6-speed and the reverse 1-speed with a small amount of change of speed change ratio between each speed change shift, the improvement of ride-in comfortableness can be promoted.

Description

DETAILED DESCRIPTION OF THE INVENTION [Industrial Field of Application] The present invention relates to a planetary gear type transmission that is incorporated into an automatic transmission used in a vehicle such as an automobile.

[Conventional technology]

Planetary gear type transmissions incorporated into automatic transmissions used in vehicles such as automobiles are generally constructed from a compound planetary gear mechanism that combines two or more sets of simple planetary gear mechanisms. They are broadly divided into the Simpson type, which is a combination of multiple single pinion planetary gear mechanisms, and the Ravigneaux type, which is a combination of a single pinion planetary gear mechanism and a double pinion planetary gear mechanism. These gears are capable of achieving three to five forward gears and a reverse gear.

This kind of planetary gear type transmission is, for example,
-27863', Japanese Patent Publication No. 50-32913, and Japanese Utility Model Application Publication No. 61-117950.

[Problems to be Solved by the Invention] In automatic transmissions, when the amount of change in the gear ratio between each gear shift becomes small due to an increase in the number of gears, the pump side of the torque converter and the turbine side The difference in rotational speed between the two sides becomes smaller, and the power transmission efficiency of the torque converter improves accordingly, leading to an improvement in the fuel economy of the vehicle. Furthermore, the smaller the relative rotational speed between the pump side and the turbine side of the torque converter during gear shifting, the more the gear shifting shock can be reduced.

From this point of view, it is required to increase the number of gears without increasing the overall gear ratio range of the transmission, that is, the amount of change between the gear ratio of the lowest speed gear and the gear ratio of the highest speed gear. ing.

However, in conventionally known planetary gear transmissions, as the number of gears increases, the number of required simple planetary gear mechanisms increases, and the number of clutches or brakes required to achieve each gear increases. This leads to an increase in the required number of frictional engagement devices such as the It is often difficult to set the gears to the required gear ratio.

In view of the above-mentioned problems, the present invention uses a combination of a few simple planetary gear mechanisms and a few frictional engagement devices to simplify the structure and achieve a desired speed ratio range without increasing the size. The object of the present invention is to provide a planetary gear type transmission having a new configuration, which can provide a large number of gears each having a predetermined gear ratio number.

[Means for Solving the Problems] According to the present invention, a single pinion type planetary gear mechanism having an input shaft, an output shaft, a sun gear gear, a planetary pinion, and a carrier, a double pinion type planetary gear mechanism having a sun gear, a ring gear, two planetary pinions, and a carrier; a third clutch that selectively connects the ring gear of the double pinion type planetary gear mechanism and the input shaft; and the single pinion. A connecting member that connects the carrier of the double pinion type planetary gear mechanism to each other, the sun gear of the single pinion type planetary gear mechanism, the sun gear of the double pinion type planetary gear mechanism, and the input. a first clutch for selectively connecting the carrier of the double pinion type planetary gear mechanism and the input shaft, and a second clutch for selectively fixing the connecting member to the transmission case. a first brake of the double pinion type planetary gear mechanism; a second brake for selectively fixing the ring gear of the double pinion type planetary gear mechanism to the transmission case; a sun gear of the single pinion type planetary gear mechanism; and a sun gear of the double pinion type planetary gear mechanism; and a third brake for selectively fixing the single pinion type planetary gear mechanism to the transmission case, and the carrier of the single pinion type planetary gear mechanism is connected to the output shaft.

[Operations and Effects of the Invention] In the planetary gear device constructed as described above, two sets of simple planetary gear mechanisms, three clutches, and three brakes are used to provide five forward speeds or forward stride and reverse speeds. A planetary gear transmission having a gear stage of You will be able to reduce the amount. A planetary gear type transmission having this speed ratio characteristic is suitable for a transmission for a passenger car where ride comfort is important.

[Example] The present invention will be described in detail below with reference to the accompanying drawings.

FIG. 1 shows a basic embodiment of a planetary gear type transmission according to the present invention.

In FIG. 1, IS indicates an input shaft, and O8 indicates an output shaft, and both are provided coaxially with each other. In the figure, 10 indicates a single-binion type planetary gear mechanism, and 20 indicates a double-binion type planetary gear mechanism, which are arranged coaxially with the input shaft Is and the output shaft O8.

Rotational power is applied to the input shaft Is from a prime mover such as an internal combustion engine (not shown) via a well-known hydraulic torque converter, electromagnetic clutch, or the like.

The output shaft O8 includes an output gear and provides rotational power to a well-known differential device (not shown).

The single pinion type planetary gear mechanism 10 includes a sun gear 11
It has a ring gear 13 concentric with the sun gear 11, a planetary pinion 12 that meshes with the sun gear 11 and the ring gear 13, and a carrier 14 that rotatably supports the planetary pinion 12. It forms a gear mechanism.

The double pinion type planetary gear mechanism 20 includes a sun gear 21, a ring gear 24 concentric with the sun gear 21, and two planetary binions 22 and 23 that mesh with each other, one meshing with the sun gear 21, and the other meshing with the ring gear 24.
and a carrier 25 rotatably supporting two planetary pinions 22 and 23, forming a general double pinion type planetary gear mechanism.

The sun gear 11 of the single pinion type planetary gear mechanism 10 and the sun gear 21 of the double pinion type planetary gear mechanism 20 are integrally connected to each other by a connecting member 26.

The carrier 25 of the double pinion type planetary gear mechanism 20 and the ring gear 13 of the single pinion type planetary gear mechanism 10 are integrally connected to each other by a connecting member 27.

The carrier 14 of the single pinion type planetary gear mechanism 10 is connected to the output shaft O8 and constitutes the only output member.

Further, the transmission device according to the present invention includes three clutches, namely, a first clutch CI, a second clutch C2, and a third clutch C3, and three brakes, namely, a first brake B1, a second brake B2, and a third clutch C3. Has three brakes &

The first clutch C1 includes a sun gear 11 of a single pinion type planetary gear mechanism 10 and a double pinion type planetary gear mechanism 2.
A connecting body with the sun gear 21 of No. 0, that is, a connecting member 26 is selectively connected to the input shaft Is in a torque transmission relationship.

The second clutch C2 is a double pinion type planetary gear mechanism 20
carrier 25 is selectively connected to the input shaft Is in a torque transmission relationship.

The third clutch G is a double vinyl-on type planetary gear mechanism 20
A ring gear 24 is selectively connected to the input shaft IS in a torque transmission relationship.

The first brake B1 is a connecting body between the ring gear 13 of the single pinion type planetary gear mechanism 10 and the carrier 25 of the double pinion type planetary gear mechanism 20, and is configured to immediately and selectively connect this to a fixed state. There is.

The second brake & is configured to selectively connect the ring gear 24 of the double pinion type planetary gear mechanism 20 to the transmission case 90 to keep it in a fixed state.

Third brake & is single-binion type planetary-wheel mechanism 1
0 sun gear 11 and double pinion type planetary gear mechanism 20
The connecting body with the sun gear 21, that is, the connecting member 26, is selectively connected to the transmission case 90 and brought into a fixed state.

In the planetary gear type transmission having the above configuration,
Since the three clutches C+, C2, Ca and the three brakes B1, &, mentioned above are engaged and released in predetermined combinations, the first gear, second gear, and third gear as deceleration gears are activated. A third gear, a fourth gear as a direct gear, a fifth gear as an increasing gear, and one reverse drive are achieved. Clutches C1, C2, C3 and brakes B1, B2, B3 in this case
The combination of engagement and release is shown in FIG. In FIG. 2, a 0 mark indicates engagement, and no mark indicates release.

In this case, the gear ratio at each gear stage is as shown in FIG. However, ρ is the ratio of the number of teeth between the sun gear 11 and the ring gear 13 of the single pinion type planetary gear mechanism 10, and ρ2 is the ratio of the number of teeth between the sun gear 21 and the ring gear 24 of the double pinion type planetary gear mechanism 20. ρ weight −0,
42, ρ2 - 0.50, the actual values of the gear ratio of each gear stage are shown in FIG.

Next, the operation at each gear stage of the planetary gear type transmission constructed as described above will be explained.

In the first gear, the first clutch C1 and the first brake B+ are engaged. At this time, as shown in FIG. 3, the power input from the input shaft Is is
4 and is transmitted to the sun gear 11 of the single pinion type planetary gear mechanism 10. In the single pinion type planetary gear mechanism 10, the ring gear 13 is fixed by engaging the first brake B1. Therefore,
The power transmitted to the sun gear 11 is taken out from the carrier 14 via the planetary pinion 12, and is transferred to the output shaft O.
8 (arrows X5, X6). The gear ratio at this time is (1+
ρth) / ρ goose. The rotation of the input shaft IS is also transmitted to the sun gear 21 of the double pinion type planetary gear mechanism 20, but the rotation of the input shaft IS is also transmitted to the sun gear 21 of the double pinion type planetary gear mechanism 20.
Since J is in the released state, the rotation on this double pinion type planetary gear mechanism 20 side is a no-load rotation. Therefore, this rotation is not involved in power transmission from the input shaft Is to the output shaft O8.

In the second gear, the second clutch C2 and the second brake & are engaged. At this time, as shown in FIG. 4, the power input from the input shaft IS is
The carrier 25 of the double pinion type planetary gear mechanism 20 is connected to the carrier 25 via the path X3, and then the carrier 2
Single pinion type planetary gear mechanism 1 connected to 5
0 ring gear 13. In the double pinion type planetary gear mechanism 20, the ring gear 24 is fixed by engaging the second brake B2, and thereby the rotational speed of each element is regulated according to the rotational speed of the carrier 25, and the sun gear 2-1 also performs a corresponding rotation. On the other hand, the ring gear 1 of the single pinion type planetary gear mechanism 10
3 rotates the sun gear 11 via the planetary pinion 12, and the rotational speed of the sun gear 11 is specified to be the same as the rotational speed of the double pinion type planetary gear mechanism 20. As a result of these, the power transmitted to the ring gear 13 is
While circulating along the path, it is taken out from the carrier 1-4 and reaches the output shaft O8 (arrows X5 and X6).
. In this case, the gear ratio is ρ2(1+ρ1)/(ρ2-ρ1
+ρ1 ・ρ2).

In the third gear, the second clutch C2 and the third brake & are engaged. At this time, as shown in FIG. 5, the power input to the input shaft Is is transmitted to the carrier 25 of the double pinion type planetary gear mechanism 20 in the order of arrows X1 to X3, and then to the single The signal is transmitted to the ring gear 13 of the pinion type planetary gear mechanism 10. Here, in the double pinion type planetary gear mechanism 20, although the sun gear 21 is fixed by the engagement of the third brake &, the third clutch a and the second brake & are in a released state. As a result of the rotation of the carrier 25, the other mouth widening elements simply rotate without load, and are not particularly involved in power transmission.

On the other hand, in the single pinion type planetary gear mechanism 10,
The sun gear 11 is fixed because the brake & is in the engaged state. Therefore, the power transmitted to the ring gear 13 is transmitted to the carrier 14 via the planetary pinion 12, and reaches the output shaft O8 (
arrows X4, X5).

The gear ratio at this time is 1+ρ1.

In the fourth gear, the second clutch C2 and the third clutch a are engaged. At this time, as shown in FIG. 6, the power input to the input shaft IS is
, via the path of X2, the double pinion type planetary gear mechanism 2
0 to the carrier 25, and the arrows X3 and X4
It is also transmitted to the ring gear 24 of the double pinion type planetary gear mechanism 20 through the path. The rotation of this ring gear 24 is the same as the rotation of the carrier 25. Therefore, each rotating element of the double pinion type planetary gear mechanism 20 rotates as a unit, and this rotation causes power to be transmitted to the single pinion type planetary gear mechanism 1 via the carrier 25 and the sun gear 21.
0 ring gear 13 and sun gear 11, and the single pinion type planetary gear mechanism 10 also rotates together (arrows X5, X6). As a result, power is extracted from the carrier 14 of the single pinion type planetary gear mechanism 10 and reaches the output shaft O8 (arrows X7, X8).

As a result of the above, the gear ratio becomes 1. This is a so-called direct connection stage.

In H, the third clutch 0 and the third brake & are engaged. At this time, as shown in Figure 7,
The power input from the input shaft Is is indicated by arrows X1 to X3.
The signal is transmitted to the ring gear 24 of the double pinion type planetary gear mechanism 20 via the path . In the double pinion type planetary gear mechanism 20, the sun gear 21 is fixed by engagement of the third brake &. As a result, the input shaft! The power transmitted from S to the ring gear 24 of the double pinion type planetary gear mechanism 20 is transmitted to the carrier 25 via the planetary pinions 22 and 23 (arrow X4), and in the single pinion type planetary gear mechanism 10. , third brake &
are engaged, so the sun gear 11 is fixed. Therefore, the power transmitted to the ring gear 13 is extracted from the carrier 14 via the planetary pinion 12 and reaches the output shaft O8 (arrows X5, X6>).

The gear ratio at this time is (1-ρ2)(1+ρI).

In reverse gear, the first clutch C1 and the second brake &
is in an engaged state. At this time, as shown in FIG. 8, the power input to the input shaft Is is
Double pinion type planetary gear mechanism 20 via route X, 3
is transmitted to the sun gear 21 of. In the double pinion type planetary gear mechanism 20, the ring gear 24 is fixed by engagement of the second brake B2. Therefore, input shaft I
From S to sun gear 2 of double pinion type planetary gear mechanism 20
The power transmitted to the planetary pinion 22
．． 23 from the carrier 25, and reaches the ring gear 13 of the single pinion type planetary gear mechanism 1o (arrow X4). In the single pinion type planetary gear mechanism 10, the sun gear 11 is the double pinion type planetary gear mechanism 20.
It rotates at the same rotational speed as the sun gear 21, but the other elements are not particularly fixed. Therefore, ring gear 1
The power transmitted to the carrier 14 circulates in the path indicated by the arrow X5 via the planetary pinion 12.
It will be taken out from the output shaft O8 (arrow X6
,X7). The gear ratio at this time is -(1-ρ 2 ) (
1+ ρ I ) / (ρ 2- ρ weight + ρ
ρ2), making it possible to transmit power for reverse deceleration.

The specific gear ratios of the first to fifth gears and the reverse gear are set to, for example, 9-0.42 and ρ2-0.50, as shown in the rightmost column of Fig. 2. A gear ratio like this can be obtained. This gear ratio is, for example,
Compared to the gear ratio of the transmission disclosed in Publication No. 17950, the overall gear ratio width between the first gear and the fifth gear is smaller, and therefore, the gear ratio can be shifted more precisely at low to medium-high speeds. Control comes into play.

In this transmission, by engaging the clutch G and the brake B1, the parallel speed ratio (ρ2 + ρ weight ・ρ
2) It is possible to separately realize the gear stage of /ρ1.

At this time, as shown in FIG. 9, the power input to the input shaft Is is transmitted to the ring gear 24 of the double pinion type planetary gear mechanism 20 in the order of arrows X1 to X3. In the double pinion type planetary gear mechanism 20, the carrier 25 is fixed by a brake B+. Therefore, the power transmitted to the ring gear 24 is transmitted to the first planetary pinion 22, the second planetary pinion 23, and further to the sun gear 21, and then to the sun gear 11 of the single-binion type planetary gear mechanism 10 (arrow X4 ,
X5). In the single-binion type planetary gear mechanism 10, the ring gear 13 is fixed by engagement of the first brake B.

Therefore, the power transmitted to the sun gear 11 is transmitted to the carrier 14 via the planetary pinion 12, and reaches the output shaft O8 (arrows X6, X7). As a result, the gear ratio (ρ2+ρ weight·ρ2)/ρ1 is realized.

Here, i set to ρ, -0,42, ρ2-0.50
, Cp2+p+ ・p2)/p+ is 1.890 (
, 1: becomes.

Therefore, depending on the power performance of the engine, this transmission can change the gear position between the second and third gears of the first to fifth gears. This can be added as a new gear to create an automatic transmission with six forward gears. (1) This transmission changes the previous third gear to the gear according to the configuration shown in FIG. 9, and the gear ratio is 3.3.
81, 2.44g, 1゜[i90.1.000.0
．． It is also possible to use an automatic transmission consisting of 71 forward speeds.

1O and 11 show specific embodiments of the planetary gear type transmission according to the present invention.

In the planetary gear type transmission shown in FIGS. 1O and 11, a first one-way clutch Fl and a second one-way clutch F2 are added to those of the above-described embodiment.

In addition, in FIG. 10, a hydraulic torque converter 80
Case 82 is shown connected to transmission case 90. Fluid torque converter 8
0 may have a general structure; this is the input shaft! S
The drive is connected to the

The first one-way clutch F1 is provided in parallel with the second brake B2, and connects the ring gear 24 of the double-binion planetary gear mechanism 20 to the transmission case 90.

The second one-way clutch F2 is the first brake B.

Single-binion type planetary gear mechanism 1
0 ring gear 13 and double binion type planetary gear mechanism 2
0 is connected to the transmission case 90 in cooperation with the first brake B1.

In the planetary gear type transmission bag device to which two one-way clutches R and R are added as described above, forward movement is achieved by engaging and disengaging the clutches and brakes according to the combination shown in FIG. A three-speed reverse gear stage can now be obtained. In Fig. 12, Q mark indicates engagement, no mark indicates disengagement, (0) mark indicates engagement when engine braking is required, and ■ mark indicates input from the input shaft Is to the transmission. The mark Q indicates that the transmission is engaged when the transmission is driven by the input shaft O8, that is, during engine braking, and the mark ◎ indicates that the transmission is engaged when the transmission is driven by the input shaft O8. 2 indicates an engagement when being driven by an input from the IS, and an engagement that does not contribute to power transmission but is engaged.

Incidentally, the gear ratio at each gear stage in this case is the same as that in the above-mentioned embodiment.

Next, the operation at each gear stage of the planetary gear type transmission constructed as described above will be explained.

In the first gear, the first clutch C1 and the first brake B are in a stopped state. Further, the second one-way clutch F2 is in a stopped state. At this time, as shown in FIG. 13, the power input from the input shaft IS is
4 and is transmitted to the sun gear 11 of the single-binion type planetary gear mechanism 10. In the single pinion type planetary gear mechanism 10, the first brake B1 is engaged,
Ring gear 13 is fixed by engagement of second one-way clutch Ft. Therefore, the power transmitted to the sun gear 11 is taken out from the carrier 14 via the planetary pinion 12 and reaches the output shaft O8 (arrow X5
,X6). Note that the rotation of the input shaft Is is also transmitted to the sun gear 21 of the double-binion type planetary gear mechanism 20, but the second brake B2, the first one-way clutch F1, and the second and third clutches c2 and C3 are in a released state. Therefore, the rotation on the double binion type planetary gear mechanism 20 side becomes a no-load rotation. Therefore, this rotation is not involved in power transmission from the input shaft Is to the output shaft O8.

Conversely, when driven by the output shaft O8, the second one-way clutch F2 idles and the ring gear 13 is unfixed, so the single-binion planetary gear mechanism 10 no longer has a fixing element, and Since the second clutch C2 is released, the power from the output shaft O8 is sufficient to cause the ring gear 13 and carrier 25 to idle, and to rotate the planetary binions 22 and 23. Therefore, at this time, the input shaft Is is not rotated by the reciprocal of the gear ratio of the first gear.

In the second gear, the second clutch C2 is engaged and the first one-way clutch F1 is engaged. At this time, as shown in FIG. 14, the power input from the input shaft IS passes through the path of arrows X'1 to
e-, t:'9 planetary gear mechanism, 20(7) carrier 2
5, and is further transmitted to the ring gear 13 of the single-binion planetary gear mechanism 10, which is connected to the carrier 25 by the connecting member 27. In the double-binion type planetary gear mechanism 20, the ring gear 24 is fixed by engaging the first one-way clutch FI, and thereby the rotational speed of each element is regulated according to the rotational speed of the carrier 2''5. The sun gear 21 also rotates accordingly.On the other hand, the ring gear 1 of the single pinion type planetary gear mechanism 1o
3 rotates the sun gear 11 via the planetary pinion 12, and the rotational speed of the sun gear 11 is specified to be the same as the rotational speed of the double-binion type planetary gear mechanism 20. As a result of these, the power transmitted to the ring gear 13 is
While circulating along the path, the carrier 14 is taken out and reaches the output shaft O8 (arrows X5, X6).

Conversely, when driven by the output shaft O8, the first one-way clutch FT idles toward the forward rotation side, and the fixed elements of both the single-binion type planetary gear mechanism 10 and the single-binion type planetary gear mechanism 20 The input shaft is gone! S is idling.

Further, when the second brake & disposed in parallel with the first one-way clutch F1 is engaged, the ring gear 24 is fixed in the forward and reverse rotation directions, and thereby the power from the output shaft O8 is transferred to the arrows X5 to X1. The input shaft Is is accelerated and reversely driven, and an engine braking effect is obtained. It becomes like this.

The third gear, the fourth gear, the fifth gear, and the reverse gear are substantially the same as those in the above-mentioned embodiment, so their explanation will be omitted. The state of the fourth gear is shown in FIG. 16, the state of the fifth gear is shown in FIG. 17, and the state of the reverse gear is shown in FIG. 18.

In addition, in reverse gear, the first one-way clutch F1 is
When the output shaft O8 is driven, the ring gear 24 is engaged together with the first brake B1, and the reaction force is received.

Also in this transmission, clutch a and brake B
1 is brought into the engaged state, the second one-way clutch F2 is engaged, and the gear ratio (ρ
A gear stage of 2+ρ weight ρ2)/ρ゛i is realized.

The state at this time is shown in FIG.

Note that even if the first brake B1 is engaged, the output shaft O8
When the ring gear 13 and the carrier 25 rotate in the normal direction, the second one-way clutch F2 becomes released, so the power from the output shaft O8 causes the ring gear 13 and the carrier 25 to idle in the normal rotation direction. However, it is not possible to increase the speed of the input shaft O8 by driving other elements.

In the above, the present invention has been described in detail with reference to five specific embodiments, but the present invention is not limited to these.Various embodiments are possible within the scope of the present invention. This will be obvious to those skilled in the art.

[Brief explanation of the drawing]

Fig. 1 is a support diagram showing a basic embodiment of the planetary gear type transmission according to the present invention, and Fig. 2 shows clutches and brakes at each gear stage of the planetary gear type transmission shown in Fig. 1. Diagrams showing the engagement states and gear ratios of FIGS. 3 to 9
The figure is a skeleton diagram showing the operating state at each gear stage of the planetary gear type transmission shown in Figure 1, and Figure 10 is a longitudinal section showing one detailed embodiment of the planetary gear type transmission according to the present invention. The top view, Figure 11 is a skeleton diagram of the planetary gear type transmission shown in Figure 10, and Figure m12 is a diagram of Figures 10 and 11.
A diagram showing the engagement state and gear ratio of the clutch and brake at each gear stage of the planetary gear type transmission shown in the figure,
FIGS. 13 to 19 are skeleton diagrams showing the operating state at each gear stage of the planetary gear type transmission shown in FIGS. 10 and 11, respectively. 10...Single binion type planetary gear mechanism, 20...
・Double pinion type planetary gear mechanism, 90... Transmission case, Summer S... Input shaft, O8... Output shaft, Q... First clutch, Ct... Second clutch, Ca... Third clutch, B+...first brake, B2...second brake, Ba...third brake, F+...first one-way clutch, F2...
・Second one-way clutch patent Applicant: Toyota Motor Corporation Representative Patent attorney
Masaki AkashiFigure 1Figure 2Figure 3Figure 4Figure 5Figure 6Figure 7 Figure 8Figure 9Figure 13Figure 14Figure 15Figure 16Figure 10Single Binion Type zuyinoIVc-*/
Planetary gear mechanism Planetary gear mechanism Figure 17 Figure 18 Figure 19

Claims (1)

[Claims] A single pinion type planetary gear mechanism having an input shaft, an output shaft, a sun gear, a ring gear, a planetary pinion, and a carrier; a double pinion type planetary gear mechanism having a sun gear, a ring gear, two planetary pinions, and a carrier; a connecting member that connects the ring gear of the pinion type planetary gear mechanism and the carrier of the double pinion type planetary gear mechanism; a sun gear of the single pinion type planetary gear mechanism; a sun gear of the double pinion type planetary gear mechanism; and the input shaft; a first clutch that selectively connects the carrier of the double pinion type planetary gear mechanism and the input shaft to each other, a ring gear of the double pinion type planetary gear mechanism and the input shaft. a third clutch selectively coupled to each other; a first brake selectively fixing the coupling member to the transmission case; and a second brake selectively fixing the ring gear of the double pinion type planetary gear mechanism to the transmission case. a third brake for selectively fixing a sun gear of the single pinion type planetary gear mechanism and a sun gear of the double pinion type planetary gear mechanism to a transmission case, the carrier of the single pinion type planetary gear mechanism; is connected to the output shaft.