Automatic transmission

The present invention relates to an automatic transmission, and particularly relates to an automatic transmission that provides six forward speeds with plural planetary gear sets, clutches and brakes.

An automatic transmission that can provide a multi-speed shift by combining plural planetary gear sets, clutches and brakes is known.

Recently, an automatic transmission with six forward speeds has been demanded for improving a fuel economy or drivability.

Herein, gear trains of such an automatic transmission need to be properly compact because a layout space of the transmission is rather limited.

For example, U.S. Pat. No. 6,176,802 has proposed such a kind of automatic transmission, in which a reduction planetary gear and a Ravigneaux-type planetary gear set (a complex-type planetary gear set with a single common carrier), which inputs a reduction rotation and a non-reduction rotation and outputs plural rotation speeds, are provided coaxially, and speed-change elements of these planetary gears are connected or disconnected with three clutches and two brakes, thereby establishing the six forward speeds.

The above-described automatic transmission can prevent the size of rotational members from being improperly large and restrain a shift shock by reducing an inertia torque, in addition to its light weight and compactness, with its structure in which two clutches are respectively disposed on both sides of the Ravigneaux-type planetary gear set.

Although the above-described automatic transmission provides the light weight, compactness and small inertia torque, it has a problem in that since an increased drive torque is transmitted to the downstream clutches and the like because of a reduction function by the reduction planetary gear, a clutch capacity would be made large enough for this increased drive torque.

Further, the planetary gear set and other members that form the downstream gear trains may need a large torque capacity as well, so the size of the gear trains would become large inevitably.

Moreover, since this automatic transmission needs to use a long pinion that extends in an axial direction for the Ravigneaux-type planetary gear set, there is a problem in that the drive torque transmission with this long pinion would cause an improper inclination of gears which generates gear noises.

Accordingly, an object of the present invention is to provide an automatic transmission for six forward speeds with plural planetary gear sets, clutches and brakes that can properly reduce the clutch capacity and restrain gear noises with its compact and light-weighted structure.

According to the present invention, there is provided an automatic transmission, comprising an input shaft, an output portion, the output portion being provided so as to be coaxial with the input shaft, a transmission case, a first single-pinion-type planetary gear set including a first sun gear, a first carrier, and a first ring gear, a second single-pinion-type planetary gear set including a second sun gear, a second carrier, and a second ring gear, a third single-pinion-type planetary gear set including a third sun gear, a third carrier, and a third ring gear, a first clutch operative to connect or disconnect the input shaft with the second sun gear, a second clutch operative to connect or disconnect the first ring gear with the third carrier, a third clutch operative to connect or disconnect the input shaft with the third sun gear, a first brake operative to connect or disconnect the third sun gear with the transmission case, and a second brake operative to connect or disconnect the third carrier with the transmission case, wherein the first sun gear and the transmission case, the first carrier and the input shaft, the second ring gear and the third carrier, and the third ring gear and the second carrier and the output portion are respectively configured to be always in a connection state, and a first speed is established by the first clutch and the second brake that are respectively selected in a connection state, a second speed is established by the first clutch and the first brake that are respectively selected in a connection state, a third speed is established by the first clutch and the third clutch that are respectively selected in a connection state, a fourth speed is established by the first clutch and the second clutch that are respectively selected in a connection state, a fifth speed is established by the second clutch and the third clutch that are respectively selected in a connection state, and a sixth speed is established by the second clutch and the first brake that are respectively selected in a connection state.

According to the above-described structure, since the first sun gear and the transmission case are always in the connection state and the first carrier and the input shaft are always in the connection state, the first single-pinion-type planetary gear set outputs the rotation with an increased rotational speed, namely it can function as a multiplying mechanism. Further, since downstream speed-change elements comprises the second and third single-pinion-type planetary gear sets, the speed-change elements can be configured without the long pinion. Thereby, the drive torque transmitted to the first and second clutches that are located downstream of the first single-pinion-type planetary gear set is not increased from the input torque, and the drive torque transmitted to the downstream gear trains is not increased either. Also, since the improper inclination of gears may not occur because of no application of the long pinion, the gear noises can be restrained.

According to an embodiment of the present invention, the third speed is configured such that a rotation of the input shaft is directly transmitted to the output portion. Thereby, the rotational speed of the fourth through sixth speeds that is higher than the third speed and may be used often in the normal driving condition can be higher than that of the input shaft. Accordingly, the drive torque acting on the gear trains of the automatic transmission can be reduced as a whole, thereby reducing a load acting on the gear trains. Thus, the durability of the gear trains can be enhanced with the light-weighted and compact structure of the automatic transmission.

According to another embodiment of the present invention, the first, second and third single-pinion-type planetary gear sets are provided on the input shaft in order from one end of the input shaft that is connected to an engine, and the output portion is provided between the first single-pinion-type planetary gear set and the second single-pinion-type planetary gear set. Thereby, in a FF (front-engine front-drive) vehicle in which its engine is disposed laterally beside the transmission, the drive torque can be outputted near the engine, namely, at a central location of the vehicle. Accordingly, a countershaft connecting with a final gear can be made properly short, thereby making the automatic transmission compact further. Particularly, since only the single-pinion-type planetary gear set is disposed between the engine and the output portion other than the torque converter, the countershaft can be made shorter. Thus, a whole part of the drive system can be made compact. Herein, a drive-force transmitting means as the “output portion” may apply a gear transmission, a chain transmission, or the like.

According to another embodiment of the present invention, the first, second and third single-pinion-type planetary gear sets are provided on the input shaft in order from an end of the input shaft that is opposite to one end of the input shaft that is connected to an engine, and the output portion is provided between the first single-pinion-type planetary gear set and the second single-pinion-type planetary gear set. Thereby, when the above-described automatic transmission is applied to the front-engine front-drive vehicle with the engine disposed laterally beside the transmission, the drive torque can be outputted relatively near the engine, namely, at the central location of the vehicle. Accordingly, the countershaft connecting with the final gear can be made properly short, thereby making the automatic transmission compact further. Thus, the whole part of the drive system can be made compact.

According to another embodiment of the present invention, the first and second clutches are disposed near the first single-pinion-type planetary gear set. Thereby, a connecting member between the first single-pinion-type planetary gear set and each of the clutches can be made properly short. Thus, there can be provided a layout that is advantageous to the compactness of the automatic transmission.

According to another embodiment of the present invention, at least either one of the first and second clutches is disposed so as not to overlap with the first single-pinion-type planetary gear set in an axial direction. Thereby, the size of either one of the clutches can be made small without any restriction by the first single-pinion-type planetary gear set. Thus, the radial-direction size of at least one of the first and second clutches can be compact, thereby making the automatic transmission compact.

According to another embodiment of the present invention, the third clutch is disposed near the third single-pinion-type planetary gear set. Thereby, a connecting member between the third single-pinion-type planetary gear set and the third clutch can be made properly short. Thus, there can be provided the layout that is advantageous to the compactness of the automatic transmission.

According to another embodiment of the present invention, the third clutch is disposed so as not to overlap with the third single-pinion-type planetary gear set in an axial direction. Thereby, the radial-direction size of the third clutch can be made small without any restriction by the third single-pinion-type planetary gear set. Thus, the radial-direction size of the third clutch can be compact, thereby making the automatic transmission compact.

According to another embodiment of the present invention, the first and second brakes are disposed side by side in an axial direction outside the second and third single-pinion-type planetary gear sets. Thereby, the brakes, the torque capacity of which may not need to be cared about, can be properly located utilizing a space outside the second and third single-pinion-type planetary gear sets. Thus, this layout can be advantageous to the compactness of the automatic transmission.

Other features, aspects, and advantages of the present invention will become apparent from the following description which refers to the accompanying drawings.

FIG. 1 is a skeleton diagram of a power train of an automatic transmission of a first embodiment according to the present invention.

FIG. 2 is a side view showing location relationships of respective shafts of the automatic transmission.

FIG. 3 is a chart showing a connection/release state of clutches and brakes.

FIG. 4A is a speed line diagram showing a speed ratio of respective speed-change elements of respective speeds, and FIG. 4B is a chart showing the number of gear tooth of respective planetary gear sets.

FIG. 5A is a chart of a gear ratio of respective speeds and a step between respective speeds, and FIG. 5B is a graph of the gear ratio.

FIG. 6A is a graph of the step between respective speeds, and FIG. 6B is a graph of a gear loss.

FIG. 7 is a chart showing a torque share ratio of respective clutches and brakes of respective speeds.

FIG. 8 is a specific sectional view of the automatic transmission.

FIG. 9 is a specific sectional view of a major part of the automatic transmission.

FIG. 10 is a skeleton diagram of a power train of an automatic transmission of a second embodiment according to the present invention.

Hereinafter, preferred embodiments of the present invention will be descried referring to the accompanying drawings.

FIG. 1 is a skeleton diagram of a power train of an automatic transmission of a first embodiment according to the present invention. FIG. 2 is a side view showing location relationships of respective shafts of the automatic transmission. This automatic transmission M is applied to a so-called FF type of power train that is disposed laterally in a vehicle as shown in FIG. 1 . The automatic transmission M is a transaxle having three shafts of an input shaft 1 , a countershaft 2 , and a drive shaft, which are provided in a vehicle width direction. On the input shaft 1 of the automatic transmission M are provided, from a side of an engine E (hereinafter, referred to as “transmission forward side”), a torque converter 4 that is equipped with a lockup clutch, a first planetary gear set G 1 that functions as a multiplying mechanism, a counter drive gear 5 that outputs a rotational drive force to the countershaft 2 , a second planetary gear set G 2 that forms a speed-change element, and a third planetary gear set G 3 that forms a speed-change element likewise.

On the countershaft 2 are provided, from a side opposite to the engine E (hereinafter, referred to as “transmission rearward side”), a counter driven gear 6 that engages with the counter drive gear 5 and a final drive gear 7 that outputs to the drive shaft 3 .

Further, on the drive shaft 3 are provided a final driven gear 8 (differential ring gear) that engages with the final drive gear 7 , a differential gear device 9 that allows differential motion between right and left wheels, and right and left drive shafts X, X that transmit the rotational force to both side wheels, not illustrated, outward from the differential gear device.

These three shafts 1 , 2 and 3 are provided in order of the input shaft 1 , the countershaft 2 and the drive shaft 3 from the vehicle forward side as shown in FIG. 2 . The countershaft 2 is disposed at a slightly higher level.

The speed-change mechanism on the input shaft 1 comprises, in addition to the above-described first, second and third planetary gear sets G 1 , G 2 , G 3 , a first clutch C 1 that is disposed on the transmission forward side of the first planetary gear set G 1 , a second clutch C 2 that is disposed on the transmission rearward side of the first planetary gear set G 1 , and a third clutch C 3 that is disposed on the transmission rearward side of the third planetary gear set G 3 , thereby transmitting the rotational force with the speed-change elements.

Further, outside the second and third planetary gear sets G 2 , G 3 are provided a first brake B 1 and a second brake B 2 that function to stop the rotation of the speed-change elements by using a transmission case Z. A one-way clutch OWC is disposed beside the second brake B 2 . The function of the one-way clutch OWC will be described later.

Each of the first, second and third planetary gear sets G 1 , G 2 , G 3 is a single-pinion-type planetary gear that includes a sun gear, a ring gear, and a carrier supporting a single pinion. Thus, the first single-pinion-type planetary gear set G 1 includes a first sun gear S 1 , a first ring gear R 1 , and a first carrier K 1 . The second single-pinion-type planetary gear set G 2 includes a second sun gear S 2 , a second ring gear R 2 , and a second carrier K 2 . The third single-pinion-type planetary gear set G 3 includes a third sun gear S 3 , a third ring gear R 3 , and a third carrier K 3 .

Next, connection relationships of the speed-change mechanism will be described. The first planetary gear set G 1 is configured to output the rotation with the increased rotational speed from the first ring gear R 1 by fixing the first sun gear S 1 to the transmission case Z (always in a connection state) and connecting the first carrier K 1 with the input shaft 1 . Thus, the first planetary gear set G 1 always functions as a multiplying mechanism.

In the second planetary gear set G 2 , the second sun gear S 2 is coupled to the first carrier K 1 via the first clutch C 1 so as to be connected or disconnected with the first carrier K 1 . And, the second carrier K 2 is always connected with the third ring gear R 3 , and the second ring gear R 2 is always connected with the third carrier K 3 . In the third planetary gear set G 3 , the third sun gear S 3 is coupled to the input shaft 1 via the third clutch C 3 so as to be connected or disconnected with the input shaft 1 . Likewise, the third carrier K 3 is always connected with the second ring gear R 2 , and the third ring gear R 3 is always connected with the second carrier K 2 . Accordingly, the second and third planetary gear sets G 2 , G 3 function as a speed-change unit that can provide four degrees of freedom in total, in which the second sun gear S 2 forms a first speed-change element (D 1 ), the second carrier K 2 third ring gear R 3 forms a second speed-change element (D 2 ), the second ring gear R 2 third carrier K 3 forms a third speed-change element (D 3 ), and the third sun gear S 3 forms a fourth speed-change element (D 4 ).

The second speed-change element (D 2 ) of the second ring gear R 2 third carrier K 3 is coupled to the first ring gear R 1 via the second clutch C 2 so as to be connected or disconnected with the first ring gear R 1 . The third speed-change element (D 3 ) of the second carrier K 2 third ring gear R 3 is always connected with the counter drive gear 5 .

Further, between the second speed-change element (D 2 ) of the second ring gear R 2 third carrier K 3 and the transmission case Z is disposed the second brake B 2 operative to stop the rotation of the third speed-change element (D 3 ). Between the fourth speed-change element (D 4 ) of the third sun S 3 and the transmission case Z is disposed the first brake B 1 operative to stop the rotation of the fourth speed-change element (D 4 ).

The automatic transmission M equipped with the above-described elements conduct a speed-change control with a control device, not illustrated, by selecting proper speeds based on a shift operation of a driver and various signals of an engine speed, a vehicle speed, a vehicle load and so on.

FIG. 3 is a chart showing a connection/release state of the clutches C 1 , C 2 , C 3 and the brakes B 1 , B 2 at respective speeds (● indicates a connection state, ∘ indicates a release state in the chart). Herein, the first clutch C 1 is also called a Low clutch because of its connection state at the first-fourth speeds. The second clutch C 2 is also called a High clutch because of its connection state at the fourth-sixth speeds. The third clutch C 3 is also called a 3/5/R clutch because of its connection state at the third and fifth speeds and the reverse. The second brake B 2 is also called a L/R brake because of its connection state at the first speed and the reverse. The first brake B 1 is also called a 2/6 brake because of its connection state at the second and sixth speeds. Also, in the present embodiment, the one-way clutch OWC is provided beside the second brake B 2 , so the second brake B 2 is not normally in the connection state at the first speed. This second brake B 2 is made in the connection state only in a specified case, for example, when an engine brake is demanded such as a manual mode or a hold mode (for this reason, an indication with ( ) is used in the chart). Herein, in a case where the one-way clutch OWC is not provided, the second brake is made in the connection state at the first speed. Accordingly, hereinafter, the control for the case where the one-way clutch OWC is not provided will be described just for convenience.

FIG. 4A is a speed line diagram showing a speed ratio of the respective speed-change elements of the respective speeds, and FIG. 4B is a chart showing the number of gear tooth of the planetary gear sets G 1 , G 2 , G 3 .

First, the speed line diagram of FIG. 4A will be descried. The axis of ordinates of this diagram indicates the respective speed-change elements of the first-third planetary gear sets G 1 , G 2 , G 3 , and a distance between these axes in a lateral direction is determined by a gear ratio. A location on the axis of ordinates indicates a speed ratio.

Specifically, three axes on the left indicates the first sun gear S 1 , the first carrier K 1 , and the first ring gear R 1 of the first planetary gear set G 1 . Four axes on the right indicates the third sun gear S 3 , the third carrier K 3 second ring gear R 2 , the third ring gear R 3 second carrier K 2 , and the second sun gear S 2 of the second and third planetary gear sets G 2 , G 3 .

The distance between the axes in the lateral direction is determined by the gear ratio, namely the number of gear tooth. For example, the distance between the first sun gear S 1 and the first carrier K 1 is determined by a gear tooth number (Zr 1 ) of the first ring gear R 1 /a gear tooth number (Zs 1 ) of the first sun gear S 1 in a case where the distance between the first carrier K 1 and the first ring gear R 1 is set to be 1 (=1.813). The distance between the third sun gear S 3 and the third carrier K 3 is determined by a gear tooth number (Zr 3 ) of the third ring gear R 3 /a gear tooth number (Zs 3 ) of the third sun gear S 3 in a case where the distance between the third carrier K 3 and the third ring gear R 3 is set to be 1 (=2.220). Likewise, the distance between the second sun gear S 2 and the second carrier K 2 is determined by a gear tooth number (Zr 2 ) of the second ring gear R 2 /a gear tooth number (Zs 2 ) of the second sun gear S 2 in a case where the distance between the second carrier K 1 and the first ring gear R 1 is set to be 1 (=1.618).

The above-described lines facilitates confirmation of the speed ratio of the respective speed-change elements. For example, in the first planetary gear set G 1 , since the first sun gear S 1 is fixed (speed ratio 0), the rotation with the same rotational direction and the increased speed (speed ratio 1.55) is outputted to the first ring gear R 1 (an intersection between a straight line interconnecting a fixed point of the first sun gear S 1 and an input point of the first carrier K 1 and an axis line of the first ring gear R 1 becomes an output point) by inputting a specified rotational speed (speed ratio 1) to the first carrier K 1 . Herein, the outputted rotational speed from the first planetary gear set G 1 can be selected by determining the output from the first ring gear R 1 (speed ratio 1.55) or the output from the first carrier K 1 (speed ratio 1).

In the second and third planetary gear sets G 2 , G 3 , there are four speed-change elements (D 1 -D 4 ) as described above. Thus, by specifying (fixing) the speed-ratio location of two speed-change elements, the speed-ratio location of the remaining two speed-change elements is determined. For example, in a case where the first clutch C 1 of the second sun gear S 2 (first speed-change element D 1 ) is made in the connection state and the second brake B 2 of the second ring gear R 2 third carrier K 3 (third speed-change element D 3 ), the second carrier K 2 third ring gear R 3 (second speed-change element D 2 ) rotates at a slower speed than the second sun gear S 2 and the third sun gear S 3 (fourth speed-change element D 4 ) rotates reversely (minus speed ratio) for idling.

The establishment of respective speeds will be described referring to FIG. 3 and FIG. 4A . The first speed is established by the first clutch C 1 and the second brake B 2 that are selected in the connection state. In this case, the rotation from the input shaft 1 by way of the first carrier K 1 is inputted to the second sun gear S 2 via the first clutch C 1 (the same rotational speed as the one of the input shaft 1 ) as shown in FIG. 1 . Since the second brake B 2 is selected in the connection state and the second ring gear R 2 is fixed to the transmission case Z, the rotational speed of the second carrier K 2 between the second sun gear S 2 and the second ring gear R 2 is greatly reduced. This reduced-speed rotation is outputted from the counter drive gear 5 .

The second speed is established by the first clutch C 1 and the first brake B 1 that are selected in the connection state. In this case, the rotation from the input shaft 1 by way of the first carrier K 1 is inputted to the second sun gear S 2 via the first clutch C 1 like the first speed. Since the first brake B 1 is selected in the connection state and the third sun gear S 3 is fixed to the transmission case Z, the rotational speed of the second carrier K 2 between the second sun gear S 2 and the third sun gear S 3 is slightly reduced. This reduced-speed rotation is outputted from the counter drive gear 5 .

The third speed is established by the first clutch C 1 and the third clutch C 3 that are selected in the connection state. In this case, the rotation that is inputted to the second sun gear S 2 from the input shaft 1 by way of the first carrier K 1 and the first clutch C 1 and the rotation that is inputted to the third sun gear S 3 from the input shaft 1 by way of the third clutch C 3 have the same rotational speed as the one of the input shaft 1 . Thus, the second sun gear S 2 and the third sun gear S 3 rotate at the same speed. Accordingly, other speed-change elements rotate at the same rotational speed, so the second carrier K 2 rotates at the same speed as well. Thereby, the rotation with the same speed as that of the input shaft 1 is outputted from the counter drive gear 5 .

The fourth speed is established by the first clutch C 1 and the second clutch C 2 that are selected in the connection state. In this case, the rotation from the input shaft 1 is also outputted to the second sun gear S 2 by way of the first carrier K 1 and the first clutch C 1 . And, the increased-speed rotation of the first ring gear R 1 is received by the second ring gear R 2 via the second clutch C 2 , so the second carrier K 2 disposed between the second sun gear S 2 and the second ring gear R 2 is increased in its rotational speed. This increased-speed rotation is outputted from the counter drive gear 5 .

The fifth speed is established by the second clutch C 2 and the third clutch C 3 that are selected in the connection state. In this case, the rotation from the input shaft 1 is outputted to the second ring gear R 2 by way of the first ring gear R 1 , the second clutch C 2 and the third carrier K 3 . Herein, since the third sun gear S 3 that is connected with the input shaft 1 via the third clutch C 3 has the same speed as the rotational speed of the input shaft 1 , the speed of the second carrier K 2 is more increased because of location relationships between the second ring gear R 2 and the third sun gear S 3 . This increased-speed rotation is outputted from the counter drive gear 5 .

The sixth speed is established by the second clutch C 2 and the first brake B 1 that are selected in the connection state. In this case, the rotation from the input shaft 1 is outputted to the second ring gear R 2 by way of the first ring gear R 1 and the second clutch C 2 . Herein, since the third sun gear S 3 is fixed to the transmission case Z further, the speed of the second carrier K 2 is much more increased because of the location relationships between the second ring gear R 2 and the third sun gear S 3 . Thus, the highest-speed rotation is outputted from the counter drive gear 5 .

The reverse is established by the third clutch C 3 and the second brake B 2 that are selected in the connection state. In this case, the rotation that is outputted to the third sun gear S 3 from the input shaft 1 via the clutch C 3 is outputted to the third ring gear R 3 with a reverse rotation because the rotation of the third carrier K 3 stops. This reverse rotation is outputted from the counter drive gear 5 .

The gear ratio of respective speeds, the steps between respective speeds, and the gear loss that are established as described above are as shown in a chart of FIG. 5A , and graphs of FIGS. 5B , 6 A, 6 B, respectively, which can provide their proper values as the automatic transmission M.

With respect to the gear ratio, for example, it is set such that the first speed is 2.6182, the second speed is 1.5026, the third speed is 1, the fourth speed is 0.7457, the fifth speed is 0.5555, the sixth speed is 0.4443, and the reverse is 2.2219, which is configured such that the higher the speed is, the less the change of the gear ratio becomes as shown in FIG. 5B . This tendency also appears in the chart of the step between the speeds, in which the step of the first and second speeds is 1.7424, the step of the second and third speeds is 1.5026, the step of the third and fourth speeds is 1.3410, the step of the fourth and fifth speeds is 1.3425, and the step of the fifth and sixth speeds is 1.2503, which is configured such that the higher the speed is, the less the change of the step between the speeds becomes as shown in FIG. 6A . By setting the gear ratio like this, the vehicle acceleration at the low speed traveling state can be improved and the comfortable vehicle riding can be attained at the high speed traveling state suppressing the shift shock properly.

Also, with respect to the gear loss (transmission loss between gears), since the single-pinion type of planetary gear sets are used, a relatively high value can be maintained as shown in FIG. 6B . Particularly, since the third speed is established by the direct connection, it provides no gear loss.

FIG. 7 is a chart showing a torque share ratio of respective clutches and brakes of respective speeds, namely to what extent the respective clutches C 1 , C 2 , C 3 and brakes B 1 , B 2 share the load of the drive torque of the input shaft 1 . Herein, “1” means that it bears all of the drive torque of the input shaft 1 , and “a value smaller than 1” means that it bears part of the drive torque of the input shaft 1 and indicates its specific extent of the bearing (share ratio). As apparent from this torque share ratio chart, the first-third clutches C 1 , C 2 , C 3 shear the transmission torque with relatively low values of 1-0.250.

This is because, as described above, the first planetary gear set G 1 provided upstream of the first and second clutches C 1 , C 2 is made function as the multiplying mechanism that transmits the rotation of the input shaft 1 to the first and second clutches C 1 , C 2 with the speed increase. Accordingly, the torque share ratio of these clutches C 1 , C 2 is relatively low. In a case where the first planetary gear set G 1 is comprised of the reduction mechanism that is shown in the above-described patent publication, the torque share ratio of these clutch C 1 and the like may increase to exceeds “1.”

Thus, the clutch capacity and the diameter size of the clutches, and the plate number of the multi-plate clutches can be reduced by lowering the torque share ratio of the clutches C 1 , C 2 , C 3 .

With respect to the brakes B 1 , B 2 , although the second brake B 2 has the value exceeding “1,” the normal traveling state in which the first speed or the reverse are used does not use this second brake B 2 so often. Accordingly, the burden of the brake may not be increased improperly.

Thus, according to the present embodiment, since the torque share ratio of the clutches C 1 , C 2 , C 3 and the brakes B 1 , B 2 can be lowered, the clutch capacity of the clutches can be reduced and thereby the clutch size can be made small. Accordingly, the compactness of the automatic transmission M with the six forward speeds can be attained effectively.

Next, the specific structure of the automatic transmission M will be further described referring to FIGS. 8 and 9 . FIG. 8 is a specific sectional view of the automatic transmission M, and FIG. 9 is a specific sectional view of a major part of the automatic transmission M. The same components as those shown in FIG. 1 will be denoted by the same reference characters, whose description is omitted.

The automatic transmission M comprises the input shaft 1 , the plural gear sets G 1 , G 2 , G 3 , the clutches C 1 , C 2 , C 3 , and the brakes B 1 , B 2 and so on, which are accommodated in the transmission case Z that comprises a converter housing Z 1 , a gear casing Z 2 and an end cover Z 3 .

A front end portion of the input shaft 1 is rotatably supported via a bearing b 1 at a cylindrical member Z 5 that is supported at a cover member Z 4 that is fixed to the gear casing Z 2 . A rear end portion of the input shaft 1 is rotatably supported via a bearing b 2 at a boss portion Z 3 a that is formed at the end cover Z 3 .

On the input shaft 1 provided the first, second and third planetary gear sets G 1 , G 2 , G 3 from the transmission forward side as described above, and the counter drive gear 5 is disposed between the first and second planetary gear sets G 1 , G 2 .

Herein, the first planetary gear set G 1 is disposed so as to overlap with the above-described cylindrical member Z 5 in the axial direction, and the first sun gear S 1 is fixed (always fixed) to the cylindrical member Z 5 via a spline connection SP 1 . Also, the first clutch C 1 is disposed just before the first planetary gear set G 1 , and the second clutch C 2 is disposed just behind the first planetary gear set G 1 .

The first clutch C 1 has a hydraulic servo Q 1 that comprises a seal plate 10 fixed to the first carrier K 1 , a piston plate 12 that is moved in the axial direction by an operating oil pressure to form a receiving chamber 11 with the seal plate 10 , and a return spring 13 that pushes back the piston plate 12 to its release position when the operating oil pressure does not act. The first clutch C 1 also has a multi-plate clutch (P 1 ) that comprises a plurality of clutch plates 16 that are disposed in the axial direction between a hub member 14 that is integral with the first carrier K 1 and a drum member 15 that is provided outside the hub member 14 .

The second clutch C 2 also has a hydraulic servo Q 2 that comprises a seal plate 20 fixed to a connection member 19 that extends from the first ring gear R 1 , a piston plate 22 that is moved in the axial direction to form a receiving chamber 21 with the connecting member 19 , and a return spring 23 pushing back the piston plate 22 to its release position. The second clutch C 2 also has a multi-plate clutch (P 2 ) that comprises a plurality of clutch plates 26 that are disposed in the axial direction between a drum portion 19 a that is formed to extend upward from the connecting member 19 and a hub member 24 that is connected with the third carrier K 3 so as to be provided inside the drum portion 19 a.

Thus, since the first and second clutches C 1 , C 2 are provided near the first planetary gear set G 1 , the rotational force can be transmitted even if the connecting members 14 , 19 are short. Specifically, the rotational force can be transmitted from the first carrier K 1 to the first clutch C 1 by just extending the first carrier K 1 outward. Also, the rotational force can be transmitted from the first ring gear R 1 to the second clutch C 2 by just providing the connecting member 19 extending in the radial direction.

Further, according to the present embodiment, the second clutch C 2 is disposed so as not to overlap with the first planetary gear set G 1 in the axial direction. Thereby, the multi-plate clutch P 2 of the second clutch C 2 can be located further inward, so the radial-direction size of the second clutch C 2 can be compact.

The counter drive gear 5 is disposed on the transmission rearward side of the second clutch C 2 . By the disposition of the counter gear drive 5 at this location, the output point of the speed-change mechanism on the input shaft 1 can be located on the transmission forward side properly. Thus, the output point of the speed-change mechanism can be provided near the central location of the vehicle (see FIG. 1 ).

A bearing member b 3 is provided on the transmission rearward side of the counter drive gear 5 . The counter drive gear 5 is firmly supported by this bearing member b 3 , thereby preventing the counter drive gear 5 from becoming improperly inclined.

Meanwhile, the second and third planetary gear sets G 2 , G 3 are disposed side by side on the transmission rearward side. The second carrier K 2 of the second planetary gear set G 2 and the third ring gear R 3 of the third planetary gear set G 3 are connected (always connected) with each other by welding, and the second ring gear R 2 of the second planetary gear set G 2 and the third carrier K 3 of the third planetary gear set G 3 are connected (always connected) with each other via a connecting portion SP 2 . Thus, the second and third planetary gear sets G 2 , G 3 are configured to function as the speed-change unit with the four degrees of freedom as described above.

The second carrier K 2 is connected (always connected) with the counter drive gear 5 via a spline connection SP 3 by way of the inside of the above-described bearing member b 3 . Thereby, the counter drive gear 5 , second carrier K 2 , and third ring gear R 3 rotate together.

The third carrier K 3 is likewise connected with the hub member 24 of the second clutch C 2 that extends by way of the inside of the bearing member b 3 via a spline connection SP 4 . The second carrier K 2 is likewise connected with the drum member 15 of the first clutch C 1 that extends by way of the inside of the bearing member b 3 via a spline connection SP 5 .

The third sun gear S 3 is supported rotatably on the input shaft 1 , and connected with the third clutch C 3 and the first brake B 1 via a drum connecting member 35 that extends outward.

The third clutch C 3 is disposed on the transmission rearward side of the third planetary gear set G 3 . The third clutch C 3 has a hydraulic servo Q 3 that comprises a seal plate 30 fixed to a hub portion 34 that is fixed to the input shaft 1 via a spline connection SP 6 , a piston plate 32 that is moved in the axial direction to form a receiving chamber 31 with the seal plate 30 , and a return spring 32 pushing back the piston plate 32 to its release position. The third clutch C 3 has a multi-plate clutch P 3 that comprises a plurality of clutch plates 36 that are disposed between the hub portion 34 and the drum connecting member 35 .

Since the third clutch C 3 is provided near the third planetary gear set G 3 , the rotational force can be transmitted even if the connecting member is short. Specifically, just by connecting the third sun gear S 3 to the drum connecting member 35 , the rotational force can be transmitted to the third sun gear S 3 from the third clutch C 3 .

Also, the third clutch C 3 is disposed so as not to overlap with the third planetary gear set G 3 in the axial direction. Thereby, the multi-plate clutch P 3 of the third clutch C 3 can be located inward further, so the radial-direction size of the third clutch C 3 can be compact.

Outside the second and third planetary gear sets G 2 , G 3 are provided the second brake B 2 , one-way clutch OWC, and first brake B 1 in order from the transmission forward side (hereinafter, see FIG. 9 ).

The second brake B 2 has a hydraulic servo Q 5 that comprises a piston plate 52 that forms a receiving chamber 51 with a partition wall Z 2 a of the gear casing Z 2 , a return spring, not illustrated, and a stopper plate, not illustrated. The second brake B 2 has a multi-plate clutch P 5 that comprises a plurality of clutch plates 56 that are disposed between a hub portion 54 that is integral with the second ring gear R 2 and an inner wall portion Z 2 b of the gear casing Z 2 .

The one-way clutch OWC has an inner lace 60 that is fixed to the hub portion 54 that is integral with the second sun gear S 2 , and an outer lace 61 that is fixed to the inner wall portion Z 2 b of the gear casing Z 2 , thereby allowing a rotation only in a one direction.

The first brake B 1 has a hydraulic servo Q 4 that comprises a piston plate 42 that forms a receiving chamber 41 with the end cover Z 3 , a return spring 43 that pushes back the piston plate 42 to its release position, and a stopper plate 40 that supports the return spring 43 . The first brake B 1 has a multi-plate clutch P 4 that comprises a plurality of clutch plates 46 that are disposed between an outside hub portion 46 a of the drum connecting member 35 and the inner wall portion Z 2 b of the gear casing Z 2 .

Thus, since the second brake B 2 , first brake B 1 and one-way clutch OWC are provided outside the second and third planetary gear sets G 2 , G 3 , a particular layout space for these members B 2 , B 1 , OWC is not necessary, so that the length of the automatic transmission M can be compact.

Particularly, there may be no need to consider the transmission torque capacity of the brakes. Accordingly, even if the brakes are provided outside the second and third planetary gear sets G 2 , G 3 , the radial-direction size of the automatic transmission M can be made short and compact as well.

Next, operations and effects of the above-described present embodiment will be described. According to the automatic transmission M of the present embodiment, by fixing the first sun gear S 1 to the gear casing Z 2 via the cylindrical member 16 and fixing the first carrier K 1 to the input shaft 1 , the first planetary gear set G 1 receiving the input rotation can be made function as the so-called multiplying mechanism. Further, since the downstream speed-change elements comprises the second and third planetary gear sets G 2 , G 3 , the speed-change elements can be configured without using the long pinion. Thereby, the drive torque transmitted to the first and second clutches C 1 , C 2 located downstream of the first planetary gear set G 1 is not increased from the input torque, and the drive torque transmitted to the downstream gear trains is not increased either. Also, since the improper inclination of gears may not occur because of no application of the long pinion, the gear noises can be restrained.

Thus, there can be provided the automatic transmission M for six forward speeds with the plural planetary gear sets G 1 , G 2 , G 3 , clutches C 1 , C 2 , C 3 and brakes B 1 , B 2 that can properly reduce the clutch capacity and restrain gear noises with its compact and the light-weighted structure.

Also, according to the present embodiment, the third speed is configured such that the rotation of the input shaft 1 is directly outputted from the counter drive gear 5 with the same rotational speed. Thereby, the rotational speed of the fourth through sixth speeds that may be used often in the normal driving condition can be higher than the one of the input shaft 1 . Accordingly, the drive torque acting on the gear trains of the automatic transmission M can be reduced as a whole, thereby reducing the load acting on the gear trains. Thus, the durability of the gear trains can be enhanced with the light-weighted and compact structure of the automatic transmission M.

Further, according to the present embodiment, the first, second and third planetary gear sets G 1 , G 2 , G 3 are provided on the input shaft 1 in order from the transmission forward side near the engine E, and the counter drive gear 5 is provided between the first and second planetary gear sets G 1 , G 2 .

Thereby, the drive torque can be outputted to the countershaft 2 at the central location of the vehicle. Accordingly, the countershaft 2 that interconnects the final driven gear 8 of the differential gear device 9 and the counter driven gear 6 can be made properly short, thereby making the automatic transmission M compact further. Particularly, since only the planetary gear set G 1 is disposed between the engine E and the counter drive gear 5 other than the torque converter 4 , the countershaft 2 can be made shorter. Thus, a whole part of the drive system can be made compact. Herein, although the present embodiment applies to the transmission system in which the drive force is transmitted between the counter drive gear 5 and the counter driven gear 6 , another system, such as a chain transmission system using a chain and sprockets, may be used.

Also, according to the present embodiment, the first and second clutches C 1 , C 2 are disposed near the first planetary gear set G 1 . Thereby, the connecting member between the first planetary gear set G 1 and each of the clutches C 1 , C 2 can be made properly short. Thus, there can be provided a layout that is advantageous to the compactness of the automatic transmission M.

Further, according to the present embodiment, the second clutch C 2 is disposed so as not to overlap with the first planetary gear set G 1 in the axial direction. Thereby, the radial-direction size of the second clutch C 2 can be made small without any restriction by the first planetary gear set G 1 . Thus, the radial-direction size of the second clutch C 2 can be compact, thereby making the automatic transmission M compact. In the case where the first clutch C 1 is disposed so as not to overlap with the first planetary gear set G 1 in the axial direction, the same effect can be provided as well.

Also, according to the present embodiment, the third clutch C 3 is disposed near the third planetary gear set G 3 . Thereby, the connecting member between the third planetary gear set G 3 and the third clutch C 3 can be made properly short. Thus, there can be provided the layout that is advantageous to the compactness of the automatic transmission M.

Further, according to the present embodiment, the third clutch C 3 is disposed so as not to overlap with the third planetary gear set G 3 in the axial direction. Thereby, the radial-direction size of the third clutch C 3 can be made small without any restriction by the third planetary gear set G 3 . Thus, the radial-direction size of the third clutch C 3 can be compact, thereby making the automatic transmission M compact.

Also, according to the present embodiment, the first and second brakes B 1 , B 2 are disposed side by side in the axial direction outside the second and third planetary gear sets G 2 , G 3 . Thereby, the brakes B 1 , B 2 , the torque capacity of which may not need to be cared about, can be properly located utilizing the space outside the second and third planetary gear sets G 2 , G 3 . Particularly, since there may be no need to keep the space for the brakes B 1 , B 2 in the axial direction, the axial-direction length (whole length) of the automatic transmission M can be short. Thus, the automatic transmission M can be made compact further.

A second embodiment will be described referring to FIG. 10 . FIG. 10 is a skeleton diagram of a power train, which corresponds to FIG. 1 . The same components will be denoted by the same reference characters, whose description is omitted.

In the present embodiment, an engine E′ and a torque converter 104 are disposed on the left side of the vehicle, and the first, second, third planetary gear sets G 1 , G 2 , G 3 are disposed in order from the right side on the input shaft 1 , like the above-described first embodiment.

Namely, an automatic transmission M′ of the present embodiment has an opposite-side layout structure to that of the first embodiment in such a manner that the third planetary gear set G 3 is located on the transmission forward side.

Thus, the counter drive gear 5 is located slightly outside in the vehicle width direction, and a counter driven gear 106 on a countershaft 102 is also located slightly outside. Herein, a disposition location of the counter driven gear 106 and a final driven gear 107 on the countershaft 102 is opposite to that of the first embodiment.

According to the present embodiment, the first, second and third planetary gear sets G 1 , G 2 , G 3 are provided on the input shaft 1 in order from the side that is far from the engine E′, and the counter drive gear 5 is provided between the first and second planetary gear sets G 1 , G 2 . Thereby, the drive torque can be outputted relatively near the engine E′, namely, at the central location of the vehicle. Accordingly, the countershaft 102 can be made properly shorter, thereby making the automatic transmission M′ compact. Thus, the whole part of the drive system can be made compact. The other operations and effects are the same as those of the first embodiment.

Herein, correspondence relationships of the structure of the present invention to the above-described embodiments are as follows. The output portion corresponds to the counter drive gear 5 . The first single-pinion-type planetary gear set corresponds to the first planetary gear set G 1 . The second single-pinion-type planetary gear set corresponds to the second planetary gear set G 2 . The third single-pinion-type planetary gear set corresponds to the third planetary gear set G 3 .

The present invention should not be limited to the above-described embodiment, but any other modifications and improvements may be applied within the scope of a spirit of the present invention.