Automatic transmission for automotive vehicles

The present description relates to an automatic transmission for automotive vehicles, particularly relates to an arrangement of gear-train of an automatic transmission.

There have been proposed various arrangements of a plurality of planetary gear sets to achieve six or more forward speed ratios. As is well known in the art, a planetary gear set consists of three rotational elements that are a sun gear, a ring gear, and a planetary carrier rotationally supporting planetary pinions, each of which engages with both the sun gear and ring gear. When a single planetary gear set is made up with only these three basic rotational elements comprising a single sun gear, a single ring gear and a single carrier, it is so called a single pinion type.

In addition to the single pinion type, there are also proposed various types of composite planetary gear sets for creating more than three rotational elements within a single planetary gear set. One example of the composite planetary gear sets is shown in U.S. Pat. No. 6,884,197, which presents a six-speed automatic transmission comprising three planetary gear sets. One of the three planetary gear sets has two sun gears, a single carrier, and a single ring gear, thereby creating four rotational elements from the one single planetary gear set.

In the '197 patent, all of the three planetary gear sets are arranged coaxially on a single input shaft. Also, an output gear is arranged coaxially on the input shaft. Therefore, a minimum number of shafts are required, so that overall size of the automatic transmission may be reduced advantageously for mounting it on an automotive vehicle, especially when the engine and the transmission are mounted laterally on the vehicle.

However, the composite planetary gear set presented in the '197 patent may have a disadvantage in noise and vibration, because the pinions may be longer for engaging with the two sun gears, and may have more tendency to tilt than the single pinion type.

Further in the '197 patent, the input shaft transmits torque to the single carrier of the planetary gear set having the two sun gears through a clutch and a connecting member arranged between the two sun gears. Because each of the sun gears needs to be connected to a frictional member or a connecting member, the clutch is forced to be located offset from the planetary gear set in the axial direction. That may cause the input shaft longer, and eventually cause the overall size of the automatic transmission greater.

Therefore, there is a need to improve an automatic transmission of the type presented in the prior art in terms of the noise or vibration and the size.

Accordingly, there is provided, in one aspect of the present description, an automatic transmission comprising a first shaft capable of receiving torque from an engine, and four planetary gear sets arranged coaxially on the first shaft and in series in an axial direction of the first shaft. Each of the four planetary gear sets has a single sun gear and a single ring gear. Further, the automatic transmission comprises an output wheel arranged coaxially on the first shaft, and capable of outputting torque transmitted from the engine through some of the four planetary gear sets. The output wheel is arranged between two of the four planetary gear sets.

By making the each of the four planetary gear sets with the single sun gear and the single ring gear, the pinions may be shorter, and the noise or vibration from the planetary gear sets in the automatic transmission may be reduced.

By arranging the output wheel between two of the four planetary gear sets, frictional elements such as brakes and clutches for selectively engaging rotational elements of the planetary gears may be overlapped with the planetary gears in the axial direction, because connecting members between the rotational elements and the frictional elements may be arranged at both sides of the each planetary gear with single sun and ring gears. Therefore, the overall length of the first shaft may be made shorter, and the overall size of the automatic transmission may be made smaller.

There is provided, in another aspect of the present invention, an automatic transmission comprising an input shaft capable of receiving torque from an engine, a transmission case rotationally supporting the input shaft, and first and second planetary gear sets arranged coaxially on the input shaft. Each of the first and second planetary gear sets has a single sun gear and a single ring gear, one of which permanently rotates together with the first shaft, and the other of which is fixed to the transmission case. The automatic transmission further comprises a third gear mechanism arranged coaxially on the input shaft between the first and second planetary gear set, and capable of receiving torque selectively through the first or second planetary gear set, and an output gear arranged coaxially on the first shaft between the first planetary gear set and the third gear mechanism and capable of outputting torque from the third gear mechanism.

Since each of the first and second planetary gear sets has, for example, the single sun gear permanently fixed to the transmission case and the single ring gear permanently rotating together with the input shaft, the first and second planetary gear sets always reduce speeds of the input shaft. Therefore, torque will be multiplied at the first and second planetary gear sets as a first stage of multiple stages, and torque load there will be minimal. This may result in a smaller radial size of the first and second planetary gear sets. Consequently, frictional element such as clutches and brakes may be overlapped with the planetary gears in the axial direction. Therefore, the overall length of the input shaft may be made shorter, and the overall size of the automatic transmission may be made smaller.

The advantages described herein will be more fully understood by reading an example of embodiments in which the above aspects are used to advantage, referred to herein as the Detailed Description, with reference to the drawings wherein:

FIG. 1 is a sectional view of transaxle including an automatic transmission for an automotive vehicle in accordance with an embodiment of the present description;

FIG. 2 is a side explanatory diagram of the transaxle of the embodiment;

FIG. 3 is a sectional view of the automatic transmission of the embodiment;

FIG. 4 is a skeleton diagram schematically showing rotational elements and frictional elements of the automatic transmission of the embodiment;

FIG. 5 is a table showing engaged frictional elements at respective speed gears;

FIG. 6 is an enlarged sectional view of an automatic transmission in accordance with another embodiment of the present description, showing a neighborhood of an output gear.

An embodiment of the present description will now be described with reference to the drawings, starting with FIG. 1 , which shows an overall construction of a transaxle TX for an automotive vehicle. The transaxle TX is coupled to an engine not shown, and is mounted laterally together with the engine on the vehicle. The transaxle TX basically consists of an automatic transmission AT, and a differential gear mechanism 4 . Then, the transaxle TX is coupled to front wheels through drive shafts 42 , and transmits torque from the engine to the front wheels.

The automatic transmission AT comprises a shifting gear mechanism 2 , and a torque converter 3 . The torque converter 3 is coupled to a crank shaft of the engine, as is known in the art. The torque converter 3 receives torque from the engine, converts or multiplies it, and transmits it to an input shaft 100 of the shifting gear mechanism 2 . The shifting gear mechanism 2 selectively takes one of speed ratios between the input shaft 100 and an output gear 200 . The output gear 200 is rotationally arranged coaxially with the input shaft 100 .

The output gear 200 engages with a driven gear 44 arranged coaxially on a counter shaft 43 , which is arranged in parallel to the input shaft 100 . The driven gear 44 is coupled to the counter shaft 43 through a spline or other coupling known in the art, and rotates together with the counter shaft 43 . Torque is transmitted from the output gear 200 to the counter shaft 43 through the driven gear 44 while the speed is reduced. A final gear 45 is also arranged coaxially on and coupled to the counter shaft 43 , and rotates together with it. The final gear engages with a ring gear 41 of the differential gear mechanism 4 , and also transmits the torque while reducing the speed. The differential gear mechanism 44 transmits the torque from the ring gear 41 to the right and left drive shafts 42 in the well known manner.

The torque converter 3 is housed within a torque converter housing 5 . The shifting gear mechanism 2 , the differential gear mechanism 4 and the counter shaft 43 are housed within a transmission case 1 . The transmission case 1 has, in general, a shifting gear chamber 12 that houses the shifting gear mechanism 2 , a differential gear mechanism chamber 13 that houses the differential gear mechanism 4 , and a counter shaft chamber 14 that houses the counter shaft arranged between the shifting gear chamber 12 and the differential gear mechanism chamber 13 . A front opening of the shifting gear chamber 12 is covered by the torque converter housing 5 , and a rear opening is covered by a rear cover 6 .

A positional relationship between the three shafts, the input shaft 100 , the counter shaft 43 and the drive shafts 42 , will now be described with reference to FIG. 2 , which shows axis centers d 1 , d 2 and d 3 of the input shaft 100 , the counter shaft 43 and the drive shafts 42 . Also FIG. 2 shows, by broken lines, the output shaft 200 that is coaxial with the input shaft 100 , the driven gear 44 and final gear 45 that are coaxial with the counter shaft 43 , and the ring gear 41 that is coaxial with the drive shafts 42 .

The three axes d 1 , d 2 and d 3 are arranged horizontally in the order of the input shaft 200 , the counter axis 43 and the drive shaft 42 . The axis d 2 of the counter shaft 43 is positioned highest among the three axes, and the axis d 3 of the drive shaft 42 is lowest. However, the relative, positional relationship between the input shaft 200 , the counter shaft 43 and the drive shaft 42 may vary in dependence on a geometric relationship of gears constituting the shifting gear mechanism 2 , gears constituting the differential gear mechanism 4 , the driven gear 44 and the final gear 45 , packaging requirement for the transaxle in an engine room of the automotive vehicle, or the like.

The shifting gear mechanism 2 will now be described with reference to FIGS. 1 , 3 and 4 . FIG. 3 shows a detailed sectional view of the shifting gear mechanism 2 in the automatic transmission AT. And, FIG. 4 schematically shows rotational elements and frictional elements of the shifting gear mechanism 2 .

As shown in FIG. 4 , the shifting gear mechanism 2 comprises first through four planetary gear sets GS 1 , GS 2 , GS 3 and GS 4 , first through three wet multi-plate clutches C 1 , C 2 and C 3 , and first and second wet multi-plate brakes B 1 and B 2 , all of which are coaxial with the input shaft 100 . The first and second planetary gear sets GS 1 and GS 2 are respectively smaller in outer diameter than the third and fourth planetary gear sets GS 3 and GS 4 , because the planetary gear sets GS 1 and GS 2 reduce speed of the input shaft 100 directly and have less torque loads than the other two, as described in greater detail below.

As shown in FIG. 1 , the transmission case 1 rotationally supports the input shaft 100 by using first and second bosses 16 a and 16 b at the both axial ends of the shifting gear chamber 12 of the transmission case 1 , through bearings and other member if any, in the manner known in the art. Also generally in the middle of the shifting gear chamber 12 in the axial direction, a support wall 15 is formed integrally with the transmission case 1 , and extends inwardly in the radial direction, supporting the input shaft 100 through a bearing and rotational members. The support wall 15 divides the shifting gear chamber 12 into a front chamber 12 a and a rear chamber 12 b . The front chamber 12 a is at a side closer to the torque converter 3 (hereinafter referred to as a front side), and a rear chamber 12 b is at a side opposite to the front side (hereinafter referred to as a rear side). The first boss 16 a is integrally formed with the transmission case 1 and extends toward the rear side, while the second boss 16 b is integrally formed with the rear cover 6 and extends toward the front side.

Only the first planetary gear set GS 1 is arranged in the front chamber 12 a , while the second through fourth planetary gear sets GS 2 , GS 3 and GS 4 are arranged in the rear chamber 12 b . Among the three planetary gears in the rear chamber 12 b , the second planetary gear set GS 2 is arranged at the end of the rear side, and the fourth planetary gear set GS 4 is arranged at the end of the front side.

Referring to FIGS. 3 and 4 , the first planetary gear set GS 1 comprises a first sun gear S 1 , a first ring gear R 1 , and a first carrier PC 1 that supports first pinions P 1 engaging with both the gears S 1 and R 1 . In other words, the first planetary gear set GS 1 comprises a single set of sun gear, ring gear and pinions engaging the both, and is called a single pinion type. Also, the second planetary gear set GS 2 is a single pinion type planetary gear set comprising a second sun gear S 2 , a second ring gear R 2 , and a second carrier that supports second pinions engaging with both the gears S 2 and R 2 . Speed reduction ratios at the first and second planetary gear sets GS 1 and GS 2 (in other words, respective ratio of numbers of teeth between the ring gears and the pinions) are different from each other.

As shown in FIG. 3 , the first sun gear S 1 is permanently fixed to an outer peripheral surface of the first boss 16 a by means of a spline or in the manner known in the art. Similarly, the second sun gear S 2 is fixed to an inner peripheral surface of the second boss 16 b by means of a spline or in the manner known in the art. On the other hand, the first ring gear R 1 is permanently connected to the input shaft 100 through a first connecting member M 1 , which has a boss fixed to the input shaft 100 in the known manner such as by welding. Similarly, the second ring gear R 2 is permanently connected to the input shaft 100 through a second connecting member M 2 , which has a boss fixed to the input shaft 100 in the known manner such as by spline.

Therefore, the ring gears R 1 and R 2 permanently rotate together with the input shaft 100 , while the sun gears S 1 and S 2 never rotate. Thereby, the first and second planetary gear sets GS 1 and GS 2 permanently reduce rotational speeds of the input shaft 100 , and output the reduced speeds respectively from the first and second carriers PC 1 and PC 2 .

The third planetary gear set GS 3 is a single pinion type planetary gear set comprising a third sun gear S 3 , a third ring gear R 3 , and a third carrier PC 3 that supports third pinions P 3 engaging with both the gears S 3 and R 3 . Also, the fourth planetary gear set GS 4 is a single pinion type planetary gear set comprising a fourth sung gear S 4 , a fourth ring gear R 4 , and a fourth carrier PC 4 that supports fourth pinions R 4 engaging with both the gears S 4 and R 4 .

Then, the third ring gear R 3 and the fourth carrier PC 4 are permanently connected by a third connecting member M 3 so that the third ring gear R 3 and the fourth carrier PC 4 may rotate together. Also, the third carrier PC 3 and fourth ring gear R 4 are (permanently) connected by means of a fourth connecting member M 4 so that the third carrier PC 3 and the fourth ring gear R 4 may rotate together. The fourth connecting member M 4 extends from the rear side of the third carrier PC 3 outwardly, and then extends outside of the third ring gear R 3 to the fourth ring gear R 4 of the fourth planetary gear set PC 4 which is arranged at the front side of the third planetary gear set PC 3 .

In other words, the third and fourth planetary gear set GS 3 and GS 4 , by being connected with each other by means of the third and fourth connecting members M 3 and M 4 , collectively create four rotational elements (the third sun gear S 3 , the third carrier PC 3 and the fourth ring gear R 4 , the third ring gear R 3 and the fourth carrier PC 4 , and the fourth sun gear S 4 ), thereby constituting a so called Simpson type planetary gear set. This particular configuration may also be called a CR-CR connection type because of the two “Carrier-Ring” connections.

The output gear 200 is arranged at the front side of the support wall 15 . A bearing 15 b is inserted and fixed to a boss 15 a formed at the inner periphery of the support wall 15 . The bearing 15 a rotationally supports the output gear 200 . The output gear 200 is connected to the fourth carrier PC 4 so that the output gear 200 and the fourth carrier PC 4 may rotate together.

The shifting gear mechanism 2 is configured, as shown in FIG. 4 , to attain six forward gears and a reverse gear by selectively engaging the five frictional elements or the three first through third clutches C 1 through C 3 and two first and second brakes B 1 and B 2 , as shown in FIG. 5 (X indicates engagement of a frictional element). Although not shown, a hydraulic control circuit is connected to actuators for the first through third clutches C 1 through C 3 and first and second brakes B 1 and B 2 , and selectively supplies a pressure to the actuators to engage the corresponding the frictional elements. The hydraulic control circuit may be a purely hydraulic controller, a electronic controller, a hydraulic and electronic controller, or any other appropriate controller known in the art.

Although described later in detail, in the present embodiment, a one-way clutch OWC is arranged in parallel to the second brake B 2 . Although normally in the first forward gear, the second brake B 2 is not engaged, the second brake B 2 is engaged only in a case such as a manual mode or a hold mode where engine brake is necessary (shown with a parenthesis in the engagement operational table). However, there may not be the one-way clutch OWC. In that case, the second brake B 2 is engaged in the first forward gear.

The first clutch C 1 engages the fourth sun gear S 4 and the first carrier PC 1 in the first through fourth gears as shown in the engagement operational table of FIG. 5 (hereinafter, the first clutch will be referred to as Low clutch). The Low clutch C 1 is arranged at the outer periphery of the first planetary gear set GS 1 in the radial direction. It comprises a clutch drum 51 , whose boss is connected to the first carrier PC 1 at its front side, a clutch hub 52 arranged at the radial inside of the clutch drum 51 , and a plurality of frictional clutch hub plates 53 and frictional clutch drum plates 54 arranged alternately in the axial direction between the clutch drum 51 and the clutch hub 52 in the radial direction.

At radial inside of the clutch drum 51 , a clutch piston 55 is arranged and defines a pressure receiving chamber with the walls of the clutch drum 51 . When a hydraulic pressure is introduced into the pressure receiving chamber, the clutch piston 55 may move toward the rear side against urging force of a return spring 57 , and push the fictional plates 53 and 54 , which press each other and transmit torque from the clutch drum 51 to the clutch hub 52 .

In the present embodiment, considering that hydraulic pressure in the pressure receiving chamber rises as it rotates due to centrifugal force, a sealing plate (balance plate) 56 is provided and defines, with the piston 55 , a centrifugal balance chamber that is next to the pressure receiving chamber. Thereby, hydraulic pressures in the centrifugal balance chamber and the pressure receiving chamber may increase equally as the rotational speed increases, so that the clutch controllability can be improved.

The clutch hub 52 extends in the axial direction to the rear side at the outer periphery of a clutch drum 61 of later described High clutch C 2 . The rear end of the clutch hub 52 couples to a disk shaped torque transmitting member 58 , which has a cylindrical boss at its inner periphery that is coupled to a fifth connecting member M 5 by a spline or the like. The fifth connecting member axially extends between the input shaft 100 and the output gear 200 as seen in the radial direction, and connects to the sun gear S 4 of the planetary gear set GS 4 . Therefore, the clutch hub 52 eventually couples to the fourth sun gear S 4 through the torque transmitting member 58 and the fifth connecting member M 5 . So, when the Low clutch C 1 is engaged, the first carrier PC 1 of the first planetary gear set GS 1 and the fourth sun gear S 4 of the fourth planetary gear set GS 4 rotate together.

The second clutch C 2 engages the third carrier PC 3 and the first ring gear R 1 (in other words, the input shaft 100 ), as shown in the engagement operational table of FIG. 5 , in the fourth through sixth gears (hereinafter, the second clutch will be referred to as a High clutch). The High clutch C 2 is arranged at the outer periphery of the first planetary gear set GS 1 in the radial direction and at the rear side of the Low clutch C 1 in the axial direction. The High clutch C 2 comprises a clutch drum 61 , whose boss is connected to the first ring gear R 1 at its rear side, a clutch hub 62 arranged at the radial inside of the clutch drum 61 , a plurality of frictional clutch hub plates 63 and frictional clutch drum plates 64 arranged alternately in the axial direction between the clutch drum 61 and the clutch hub 62 in the radial direction. The clutch hub 62 has a cylindrical boss at its radial inner periphery, which axially extends to the rear side between the input shaft 100 and the output gear 200 in the radial direction, and eventually connects to the third carrier PC 3 . Accordingly, when the above High clutch C 2 is engaged, the first ring gear R 1 of the third carrier PC 3 of the third planetary gear set GS 3 will rotate together.

The High clutch C 2 has, at its radial inside, a clutch piston 65 , which defines a pressure receiving chamber with walls of the first connecting member M 1 that connects the first ring gear R 1 to the input shaft 100 . The clutch piston 65 may move, depending on operational hydraulic pressure supplied to the pressure receiving chamber, against urging force of a return spring 67 , toward the rear side (in the left direction in FIG. 3 ), press the clutch plates 63 and 64 to each other, and engage together the clutch drum 61 and the clutch hub 62 . Considering the centrifugal force applied on the operational oil in this pressure receiving chamber as well, a sealing plate 66 is arranged to define a centrifugal balance chamber next to the pressure receiving chamber.

As described above, the two clutches C 1 and C 2 are housed in the front chamber 12 a of the shifting gear chamber 12 of the transmission case 1 . The Low clutch C 1 which selectively transmits torque from the input shaft 100 to the sun gear S 4 of the fourth planetary gear set is arranged at the front side of the High clutch C 2 which selectively transmits torque from the input shaft 100 to the planetary carrier PC 3 of the third planetary gear set GS 3 and the sun gear S 4 of the fourth planetary gear set PC 4 . The Low clutch C 1 and the High clutch C 2 are arranged aligned in the input shaft direction, and the clutch piston 55 of the Low clutch C 1 and the clutch piston 65 of the High clutch C 2 are respectively arranged at the front and rear sides of the first planetary gear set GS 1 in the axial direction. Therefore, these rotational parts are housed in the front chamber 12 a with higher space efficiency, and the members connecting between these rotational parts can be can be made shorter.

On the other hand, in the rear chamber 12 b of the shifting gear chamber 12 of the transmission case 1 , three frictional elements, i.e. the first clutch C 3 and the first and second brakes B 1 and B 2 , are housed.

The third clutch C 3 may engage the second carrier PC 2 and the third sun gear S 3 , as shown in the engagement operational table of FIG. 5 , in the third gear, the fifth gear and the reverse gear (hereinafter, the third clutch will be called a 3/5/R clutch). The 3/5/R clutch C 3 has a clutch drum 71 which is arranged at the outer periphery of the second planetary gear set GS 2 , has a boss connected to the second carrier PC 2 , a clutch hub 72 which is arranged at the inner peripheral side of the clutch drum 71 in the radial direction and has a boss connected to the third sun gear S 3 , and a plurality of frictional clutch hub plates 73 and frictional clutch drum plates 74 arranged alternately in the axial direction between the clutch drum 51 and the clutch hub 52 in the radial direction. Accordingly, when the 3/5/R clutch is engaged, the second carrier PC 2 of the second planetary gear set GS 2 and the third sun gear S 3 of the third planetary gear set GS 3 will rotate together.

The 3/5/R clutch C 3 has, at its radial inside, a clutch piston 75 , which defines a pressure receiving chamber with walls of the clutch drum 71 . The clutch piston 75 may move, depending on operational hydraulic pressure supplied to the pressure receiving chamber, against urging force of a return spring 77 , toward the front side (in the left direction in FIG. 3 ), press the clutch plates 73 and 74 to each other, and engage together the clutch drum 71 and the clutch hub 72 . Considering the centrifugal force applied on the operational oil in this pressure receiving chamber as well, a sealing plate 76 is arranged to define a centrifugal balance chamber with the clutch piston 75 .

The first brake B 1 engages the third sun gear S 3 to the transmission case 1 , and stops the rotation of the third sun gear S 3 , as shown in the engagement operational table of FIG. 5 , in the second and sixth gears (hereinafter, the first brake will be referred to as a 2/6 brake). The 2/6 brake B 1 has a brake hub 82 that is arranged at between the 3/5/R clutch C 3 and the transmission case 1 in the radial direction, and has a boss connected to the clutch hub 72 of the 3/5/R clutch C 3 , and a plurality of frictional brake hub plates 83 and frictional brake case plates 84 arranged alternately in the axial direction between the inner surface of the transmission case and the brake hub 82 in the radial direction. Since the brake hub 82 connects to the clutch hub 72 of the 3/5/R clutch C 3 , the brake hub 82 eventually connects to the third sun gear S 3 . Accordingly, the 2/6 brake B 1 is engaged, the sun gear S 3 of the third planetary gear set GS 3 does not rotate.

A brake piston 85 is inserted into a concave formed in the rear cover 6 of the transmission case 1 , and defines a pressure receiving chamber with that concave. The brake piston 85 may move, depending on a hydraulic pressure supplied to the pressure receiving chamber, against urging force of a return spring 87 , toward the front side (in the right direction in FIG. 3 ), press the brake plates 83 and 84 to each other, and engage together the brake hub 82 and the transmission case 1 .

The brake piston 85 of the 2/6 brake B 1 is arranged at the outer side of the clutch piston 75 of the 3/5/R clutch C 3 in the radial direction, and the both pistons 75 and 85 are overlapped with each other in the axial direction. Therefore, the both pistons 75 and 85 can be arranged closer to the torque converter 3 in the axial direction, and the overall length of the automatic transmission AT can be made shorter.

The second brake B 2 engages the fourth connecting member M 4 , which connects the third carrier PC 3 and the fourth ring gear R 4 , to the transmission case 1 , and selectively stops the rotation of the fourth connecting member M 4 and the fourth ring gear R 4 , as shown in the engagement operational table of FIG. 5 , in the first and reverse gears (hereafter, the second brake B 2 will be called a L/R brake). As described above, the L/R brake B 2 is engaged in the first gear only when engine brake is needed, such as in a manual mode, a hold mode, or the like.

The L/R brake B 2 is arranged at the front side of the 2/6 brake B 1 , and an outer side of the third and fourth planetary gear sets GS 3 and GS 4 . The fourth connecting member M 4 constitutes a brake hub of the L/R brake B 2 . A plurality of frictional brake hub plate 93 and frictional brake case plate 94 are arranged alternately in the axial direction and between the fourth connecting member M 4 and the transmission case in the radial direction. Accordingly, when the L/R brake B 2 is engaged, the third carrier PC 3 of the third planetary gear set GS 3 and the ring gear R 4 of the fourth planetary gear set GS 4 do not rotate.

A brake piston 95 is inserted into a concave formed in the support wall 15 at its rear side to define a pressure receiving chamber. The brake piston 95 may move, depending on a hydraulic pressure supplied to the pressure receiving chamber, against urging force of a return spring 97 , toward the rear side (left side of FIG. 3 ), press the brake plates 93 and 94 to each other, and engage together the fourth connecting member M 4 and the transmission case 1 .

The return spring 97 is arranged in a concave groove 17 formed at the inner surface of the transmission case 1 , and thereby the return spring 97 is positioned at an outer radial side of the brake plates 93 and 94 .

At the rear side of the L/R brake B 2 , a one-way clutch OWC is arranged between the transmission case 1 and the fourth connecting member M 4 . The one-way clutch OWC prevents the fourth connecting member M 4 from rotating in one direction.

The one way clutch OWC is supported by a support member 18 that is attached and fixed to the inner surface of the transmission case 1 . The support member 18 also functions as a receiver for the spring force of the return spring 97 of the L/R brake B 2 , and as a retainer plate to for receiving the pressing force of the brake plates 93 and 94 caused by the brake piston 95 .

The automatic transmission AT of the present embodiment has the first through fourth planetary gear sets GS 1 through GS 4 of the single pinion type, and configured to selectively activate the three first through third clutches C 1 through C 3 and two first and second brakes B 1 and B 2 to attain six forward gears and a reverse gear. In other words, the automatic transmission AT does not include a compound type planetary gear set such as the Ravigneaux type, does not need a planetary gear set of the double sun gear type or double ring gear type, and further needs only small number, actually five or less, of frictional elements that engage rotational elements with each other, for which, it is advantageous in reduction of cost and weight, reduction of noise or vibration.

As described above, the first and second planetary gear sets GS 1 and GS 2 permanently reduce the speed of the input shaft 100 , and torque from the input shaft is permanently input to the planetary gear sets GS 1 and GS 2 through no releasable frictional elements and multiplied there. Therefore, torque load there is a minimum because the planetary gear sets GS 1 and GS 2 perform a first stage of multiple torque multiplication stages, and torque fluctuation there is a minimum because of the permanent torque transmission. Consequently, the first and second planetary gear set can have a smaller size. Thereby, the Low and High clutches C 1 and C 2 can be arranged at the outer side of and partly overlapped with the first planetary gear set GS 1 in the axial direction, and the 3/5/R clutch C 3 and the 2/6 brake B 1 can be arranged at the outer side of and overlapped with the second planetary gear set GS 2 .

Further, by arranging the first planetary gear set GS 1 at the front side, the second permanent reduction gear set GS 2 at the rear side, the Simpson type planetary gear set consisting of the relatively larger two sets of third and fourth planetary gear sets GS 3 and GS 4 between the first and second planetary gear sets GS 1 and GS 2 , and the output gear 200 that is connected to the fourth carrier PC 4 between the Simpson type planetary gear set and the first planetary gear set GS 1 , the connecting members and the like M 1 through M 5 , 51 , 52 , 58 , 61 , 62 , 71 , 72 and 82 can be made shorter, and the automatic transmission AT can be made smaller.

Also, by arranging the Low and High clutches C 1 and C 2 at the outer side of the first planetary gear set GS 1 , the overall length of the transmission can be made shorter to construct it compact, as well as to make it further appropriate for mounting on front wheel drive vehicles. At this time, since the first planetary gear set GS 1 is made compact, the automatic transmission AT is prevented from being larger in the radial direction even when the clutches C 1 and C 2 are arranged at the outer side of the first planetary gear set GS 1 .

The 2/6 brake B 1 and the L/R brake B 2 , which are relatively large in diameter especially from a viewpoint of securing brake capacity, are housed in the rear chamber 12 b . Only the clutches (Low and High clutches C 1 and C 2 ), which do not need so large diameter, are housed in the front chamber 12 a , in which the first planetary gear set GS 1 is housed, and arranged side by side in the axial direction. Therefore, the final gear 45 can be arranged overlapped with the two clutches C 1 and C 2 in the axial direction, and eventually the differential gear mechanism 4 can be too. Consequently, the differential gear mechanism 4 can be arranged closer to the torque converter 3 or the vehicle center line.

Further by arranging the first planetary gear set GS 1 in the substantially middle position in the axial direction in the front chamber and the clutch pistons 55 and 65 of the Low and High clutches C 1 and C 2 in spaces of the both sides of the first planetary gear set GS 1 in the axial direction, the first planetary gear set GS 1 and the Low and High clutches C 1 and C 2 can be arranged with great space efficiency.

Additionally by arranging the Low clutch C 1 that is connected to the output of the first planetary gear set GS 1 at its the front side, and the High clutch C 2 that is connected to the input of the first planetary gear set GS 1 at its rear side, the members 51 , 52 , 58 , 61 and 62 that make input/output with the Low and High clutches C 1 and C 2 can be arranged in shorter length, and then in thickness, so the automatic transmission AT can be made smaller.

Since the front chamber 12 a can be made smaller in the radial direction as described above, the differential gear mechanism 4 can be closer to the front chamber 12 a in the vehicle longitudinal direction (see FIG. 2 ), and the overall transaxle TX including the differential gear mechanism 4 can be smaller, and the differential gear mechanism 4 can be arranged closer to the vehicle center line (see FIG. 1 ). This is advantageous on that the both drive shafts 42 make their folding angles to be moderate when the front wheels are steered. Also, the ring gear 41 of the differential gear mechanism 4 can be made greater in diameter, and it will be preferable on that the design freedom of speed reduction ratio is increased without decreasing compactness.

By arranging 3/5/R clutch C 3 and the 2/6 brake B 1 at the outer side of the second planetary gear set GS 2 in the radial direction and overlapped in the axial direction, these parts can be housed in smaller space in the axial and radial directions because the second planetary gear set GS 2 is made smaller as described above. Therefore, the rear cover 6 , the transmission case 1 or the like can be made smaller while their rigidity can be maintained.

Further, since the output gear 100 is arranged at the rear side of the first planetary gear set GS 1 , the driven gear 44 of relatively greater diameter can be arranged offset from the first planetary gear set GS 1 in the vehicle width direction, and the counter shaft 43 can be arranged closer to the shifting gear mechanism 2 in the generally vertical direction (see FIG. 2 ). In other words, the transaxle TX can be made smaller in the vertical direction. And, since the length of the counter shaft 43 can be made shorter by arranging the driven gear 44 closer to the final gear 45 , the transaxle can be made smaller in the vehicle width direction without causing unnecessary mechanical loss or weight increase (see FIG. 1 ).

In the above embodiment, the output gear 100 is rotationally supported by means of the bearing 15 b inserted into the boss 15 a of the support wall 15 , arranged at the front side of the support wall 15 , and housed in the front chamber 12 a . Therefore, the driven gear 44 can be arranged closer to the torque converter in the axial direction, and eventually the counter shaft 43 can be made shorter.

Alternatively, as shown in FIG. 6 , a bearing 15 b ′ may be outfit on the boss 15 a ′ of the support wall 15 ′, and the output gear 200 may be supported by means of the bearing 15 b ′, and housed in a concave groove formed in the support wall 15 ′ at its rear side. In this construction, the output gear 200 is not required to pass through the support wall 15 ′ to connect to the planetary carrier PC 4 of the fourth planetary gear set GS 4 , and the input shaft 100 can be made shorter.

It is needless to say that the invention is not limited to the illustrated embodiments and that various improvements and alternative designs are possible without departing from the substance of the invention as claimed in the attached claims. For example, the transaxle TX may be mounted on rear wheel drive vehicles.