Abstract:
An automatic transmission, which has at least six selectable forward speeds, includes a reduction planetary gearset that reduces speed of rotation input from an engine, a plurality of planetary gearsets disposed behind the reduction planetary gearset, a plurality of clutches that are disposed radially beyond the plurality of planetary gearsets, and a plurality of brakes that are disposed radially beyond the plurality of clutches. The plurality of planetary gearsets includes a ring gear as an input member that inputs reduced rotation from the reduction planetary gearset. Individually, planetary gearsets include a single set of pinion gears. The forward speeds of the transmission are selectable through a combination of engagement and disengagement of the plurality of clutches and the plurality of brakes. Each brake overlaps with a corresponding clutch in the axial direction.

Description:
BACKGROUND OF THE INVENTION 
     The present invention relates to an automatic transmission. In particular, it relates to technology which aims to increase the compactness of an automatic transmission comprising an input member, a plurality of planetary gearsets, three clutches, two brakes, and an output member, with at least six forward gears and reverse being attainable through appropriate engagement and disengagement of the three clutches and two brakes. 
     The recent trend is to make automatic transmissions which are more stepped or graduated to provide slightly more forward gear ratios that can be selected in an aim for improvement in drivability and fuel efficiency. Japanese Patent Provisional Application 2000-55152 discloses an automatic transmission where six forward speeds and reverse are selectable and employs a double-pinion type Ravigneaux compound planetary gear train which has two sets of planet-pinions, each set being meshed with a different sun gear, and will now be discussed with reference to  FIG. 9 . The automatic transmission disclosed therein comprises a first clutch C 10  and a second clutch C 20  which are selectably engageable to direct rotation that has been reduced in speed. The automatic transmission also comprises a planetary gear train G, which as has been stated is a Ravigneaux compound type, and it is common to dispose first clutch C 10  and second clutch C 20  behind and around planetary gear train G, and to dispose a first brake B 10 , in this case a band brake as shown in  FIG. 9 , about an outer circumference of first clutch C 10  and second clutch C 20 , and to dispose a second brake B 20  in a single row with first clutch C 10  and second clutch C 20 . 
     However, according to the related art, first clutch C 10 , second clutch C 20 , and second brake B 20 , which is a clutch-type brake, are disposed in a row in the axial direction of the transmission assembly around an outer periphery of planetary gear train G, and this leads to a problem of the axial length of the transmission assembly being large. Also, one of the planetary gearsets constituting planetary gear train G is a double-pinion planetary gearset, and rotation having an increased torque after being reduced in speed by the reduction planetary gearset is input to planetary gear train G from a first sun gear S 10  and a second sun gear S 20  thereof, meaning that an outer diameter of planetary gear train G must be large, and consequently the transmission assembly becomes undesirably large overall. 
     It is therefore an object of the present invention to provide an automatic transmission which is more compact in both the axial and radial dimensions. 
     SUMMARY 
     An aspect of the present invention resides in an automatic transmission which has at least six selectable forward speeds, the automatic transmission comprising a reduction planetary gearset which reduces speed of rotation input from an engine, a plurality of planetary gearsets disposed behind the reduction planetary gearset, the plurality of planetary gearsets comprising a ring gear as an input member which inputs reduced rotation from the reduction planetary gearset, individual planetary gearsets respectively comprising a single set of pinion gears, a plurality of clutches which are disposed radially beyond the plurality of planetary gearsets, and a plurality of brakes which are disposed radially beyond the plurality of clutches, the forward speeds of the transmission being selectable through a combination of engagement and disengagement of the plurality of clutches and the plurality of brakes, each brake overlapping with a corresponding clutch in the axial direction. 
     Another aspect of the present invention resides in an automatic transmission which has at least six selectable forward speeds, the automatic transmission comprising a reduction planetary gearset, a rear planetary gear train disposed behind the reduction planetary gearset, the rear planetary gear train comprising a first rear planetary gearset which is disposed behind the reduction planetary gearset to receive a reduced rotation therefrom, the first rear planetary gearset comprising a sun gear, a single set of pinion gears meshing with the sun gear, and a ring gear meshing with the single set of pinion gears, the ring gear being an input member which inputs the reduced rotation from the reduction planetary gearset, and a second rear planetary gearset which is disposed behind the first rear planetary gearset, the second rear planetary gearset comprising a single set of pinion gears, a plurality of clutches disposed in a row around the rear planetary gear train, and a plurality of brakes disposed in a row around the plurality of clutches, the plurality of brakes comprising a first brake which overlaps in the axial direction with a first clutch of the plurality of clutches, and a second brake which overlaps in the axial direction with a second clutch of the plurality of clutches. 
     A further aspect of the present invention resides in an automatic transmission comprising an input member which inputs an engine rotation, a planetary gear train to receive the engine rotation from the input member, the planetary gear train comprising a first planetary gearset acting as a reduction planetary gearset which inputs the engine rotation from the input member, a second planetary gearset disposed behind the first planetary gearset, the second planetary gearset comprising a sun gear, planetary pinions which mesh with the sun gear, a pinion carrier which supports the planetary pinions to be freely rotatable, and a ring gear which meshes with the planetary pinions, the ring gear acting as an input member which inputs reduced rotation from the first planetary gearset, and a third planetary gearset disposed behind the second planetary gearset, the third planetary gearset comprising two sun gears, common planetary pinions which mesh with the two sun gears, a pinion carrier which supports the planetary pinions to be freely rotatable, and a ring gear which meshes with the planetary pinions, an output member disposed coaxially with the input member, the output member receiving a rotation from the planetary gear train, and three clutches and two brakes, at least six forward speeds and reverse speed being selectable by selective engagement and disengagement of the three clutches and the two brakes, two clutches of the three clutches being disposed around the planetary gear train, the two brakes being disposed around the two clutches, one clutch of the two clutches and one brake of the two brakes overlapping in the axial direction at least partially, the other clutch of the two clutches and the other brake of the two brakes overlapping in the axial direction at least partially. 
     The other objects and features of this invention will become understood from the following description with reference to the accompanying drawings. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a skeleton diagram of an embodiment of the present invention. 
         FIG. 2  is a chart diagram of engagement logic showing a relationship between selectable gears and engagement of transmission friction elements of the same automatic transmission. 
         FIG. 3A  is a skeleton diagram showing power flow of torque in a first forward gear in the same automatic transmission. 
         FIG. 3B  is a skeleton diagram showing power flow of torque in a second forward gear in the same automatic transmission. 
         FIG. 3C  is a skeleton diagram showing power flow of torque in a third forward gear in the same automatic transmission. 
         FIG. 4A  is a skeleton diagram showing power flow of torque in a fourth forward gear in the same automatic transmission. 
         FIG. 4B  is a skeleton diagram showing power flow of torque in a fifth forward gear in the same automatic transmission. 
         FIG. 4C  is a skeleton diagram showing power flow of torque in a sixth forward gear in the same automatic transmission. 
         FIG. 5  is a skeleton diagram showing power flow of torque in a reverse gear in the same automatic transmission. 
         FIG. 6  is a cross-sectional view showing the same automatic transmission. 
         FIG. 7  is a skeleton diagram showing physical location of internal elements of the same automatic transmission. 
         FIG. 8  is an enlarged cross-sectional diagram showing detail of the same automatic transmission. 
         FIG. 9  is a cross-sectional view of an automatic transmission using a Ravigneaux compound planetary gear train according to the related art. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     Referring to  FIGS. 1 through 8 , there is discussed an embodiment of an automatic transmission in accordance with the present invention. 
     Referring to  FIG. 1 , an automatic transmission according to the present invention for use in an automobile comprises a first planetary gearset G 1 , a second planetary gearset G 2 , a third planetary gearset G 3 , a first connecting member M 1 , a second connecting member M 2 , a first clutch C 1 , a second clutch C 2 , a third clutch C 3 , a first brake B 1 , a second brake B 2 , an input member Input, being an input shaft  1 , and an output member Output, being an output gear  2 . The disposition and relationships of these elements will now be explained. 
     The embodiment of the automatic transmission comprises, starting from the left of  FIG. 1  near input member Input, first planetary gearset G 1  as a reduction mechanism comprised of a single-pinion planetary gearset having a single set of planet-pinions, second planetary gearset G 2  comprised of a single-pinion planetary gearset also having a single set of planet-pinions, and third planetary gearset G 3  comprised of a double-sun-gear planetary gearset having two sun gears, all being disposed coaxially. First planetary gearset G 1  functions as a reduction planetary gearset, and second planetary gearset G 2  and third planetary gearset G 3  constitute a compound planetary gear train located in a rear portion of the transmission. 
     First planetary gearset G 1 , which is a single-pinion planetary gearset serving as a reduction planetary gearset, is comprised of a first sun gear S 1 , a first ring gear R 1 , first planet-pinions P 1  which mesh with first sun gear S 1  and first ring gear R 1 , and a first carrier PC 1  which supports first planet-pinions P 1  to be freely rotatable. Second planetary gearset G 2  is a single-pinion planetary gearset comprised of a second sun gear S 2 , a second ring gear R 2 , second planet-pinions P 2  which mesh with second sun gear S 2  and second ring gear R 2 , and a second carrier PC 2  which supports second planet-pinions P 2  to be freely rotatable. 
     Third planetary gearset G 3  is a double-sun-gear planetary gearset comprised of a third sun gear S 3  and a fourth sun gear S 4 , third planet-pinions P 3  which mesh both with third sun gear S 3  and fourth sun gear S 4 , a third carrier PC 3  which supports third planet-pinions P 3  to be freely rotatable, and a third ring gear R 3  which meshes with third planet-pinions P 3 . Third sun gear S 3  and fourth sun gear S 4  are disposed such that third sun gear S 3  is nearer input member Input than fourth sun gear S 4 . Third sun gear S 3  and fourth sun gear S 4  are disposed coaxially, but it is not necessary for third sun gear S 3  and fourth sun gear S 4  to have the same number of teeth. An automatic transmission according to the present invention also comprises a center member CM and an outer member OM joined to third carrier PC 3 , center member CM extending radially inward toward the axis from between third sun gear S 3  and fourth sun gear S 4 , and outer member OM extending radially outward away from the axis. Outer member OM is uniquely disposed, and this will be discussed in more detail. Further, center member CM extends radially inward toward the axis so that it passes through space existing between individual planet-pinions of third planet-pinions P 3 . 
     Input member Input comprises input shaft  1 , and input shaft  1  is joined to first ring gear R 1 , and is coupled to an engine (not shown) through a torque converter (not shown), so that engine rotation is input into first ring gear R 1  from input shaft  1 . Output member Output comprises output gear  2 , and is joined coaxially to second connecting member M 2  which serves to join second carrier PC 2  and third ring gear R 3 . Output rotation from the transmission is transmitted, for example, from output gear  2  to counter gear  30  shown in  FIG. 6 , then on to a final gear and differential gear apparatus (neither of which shown) to the drive wheels of an automobile. Further, first connecting member M 1  serves to join second sun gear S 2  and third sun gear S 3  to form a single integral body. 
     First sun gear S 1  of reduction planetary gearset G 1  is permanently fixed to transmission case  3 , and first carrier PC 1  is appropriately joinable to second ring gear R 2  by first clutch C 1  and is appropriately joinable to second sun gear S 2  by second clutch C 2 . Center member CM of third carrier PC 3  is appropriately joinable to input shaft  1  by third clutch C 3 , and therefore third clutch C 3  serves as a direct clutch to transmit input rotation directly to the compound planetary gear train comprised of second planetary gearset G 2  and third planetary gearset G 3 . Outer member OM of third carrier PC 3  of third planetary gearset G 3 , which is a double-sun-gear planetary gearset, is appropriately joinable to transmission case  3  by first brake B 1  so that third carrier PC 3  is made appropriately fixable, and fourth sun gear S 4  is made appropriately fixable to transmission case  3  by second brake B 2 . 
     It is possible to select gears, i.e., forward speeds 1st through 6th and reverse, with the gearshift assembly of the present invention through the corresponding combinations of friction elements first clutch C 1 , second clutch C 2 , third-clutch C 3 , first brake B 1 , and second brake B 2 , as shown by the engagement logic table in  FIG. 2 , where engagement is represented by a circle mark and disengagement by being unmarked. A control valve body for controlling gear shift (not shown) is connected to first clutch C 1 , second clutch C 2 , and third clutch C 3  as well as first brake B 1  and second brake B 2  to realize the engagement logic. A hydraulic type, an electronic type, or a combination type which combines these two types are employable as a control valve body for control of gear shift. 
     Below, operation of an automatic transmission according to the present invention will be discussed with reference to  FIGS. 2 through 5 . In  FIGS. 3 through 5 , the power flow of torque through first clutch C 1 , second clutch C 2 , third clutch C 3 , first brake B 1 , second brake B 2 , center member CM, and outer member OM is indicated by thick lines, and gears which participate in torque transmission are indicated by hatching. 
     (First gear) As shown by  FIG. 2 , first gear is achieved through engagement of first clutch C 1  and first brake B 1 . In first gear, reduced rotation from first planetary gearset G 1  is input into second ring gear R 2  of second planetary gearset G 2  by engagement of first clutch C 1 . At the same time, third carrier PC 3  of third planetary gearset G 3  is fixed to transmission case  3  by engagement of first brake B 1 , so rotation becomes a reverse-direction reduced rotation. Rotation of third sun gear S 3  is then transmitted to second sun gear S 2  of second planetary gearset G 2  via first connecting member M 1 . Thus, at second planetary gearset G 2 , a normal-direction reduced rotation is input from second ring gear R 2 , and a reverse-direction reduced rotation is input from second sun gear  52 , and as a result, a rotation which is a further reduced rotation from second ring gear R 2  is output to output gear  2  via second connecting member M 2  from second carrier PC 2 . 
     The power flow in first gear is as shown in  FIG. 3A . Torque acts through first clutch C 1 , first brake B 1 , first connecting member M 1 , and second connecting member M 2 , shown in bold lines, and first planetary gearset G 1 , second planetary gearset G 2 , and third planetary gearset G 3  not including fourth sun gear S 4 , shown in hatching. Therefore, in first gear, all planetary gearsets are involved in transmission of torque, that is, first planetary gearset G 1 , as well as second planetary gearset G 2  and third planetary gearset G 3  which make up the compound planetary gear train located in the rear portion of the transmission. 
     (Second gear) As shown by  FIG. 2 , in second gear, first brake B 1  which was engaged in first gear is disengaged, and second brake B 2  is engaged instead. Therefore, second gear is achieved through engagement of first clutch C 1  and second brake B 2 . In second gear, reduced rotation from first planetary gearset G 1  is input into second ring gear R 2  of second planetary gearset G 2  by engagement of first clutch C 1 . At the same time, fourth sun gear S 4  of third planetary gearset G 3  is fixed to transmission case  3  by engagement of second brake B 2 , so third sun gear S 3  connected to fourth sun gear S 4  by third planet-pinions P 3  is fixed. Second sun gear S 2  which is joined to third sun gear S 3  by first connecting member M 1  is then fixed to transmission case  3 . Thus, at second planetary gearset G 2 , normal-direction reduced rotation is input from second ring gear R 2 , and second sun gear S 2  is fixed, and as a result, a reduced rotation from second ring gear R 2  which has been further reduced is output to output gear  2  through second connecting member from second carrier PC 2 . This rotation is faster than the rotation in first gear. 
     The power flow in second gear is as shown in  FIG. 3B . Torque acts through first clutch C 1 , second brake B 2 , first connecting member M 1 , and second connecting member M 2 , shown in bold lines, and first planetary gearset G 1  and second planetary gearset G 2 , shown in hatching. Further, regarding third planetary gearset G 3 , third planet-pinions P 3 , which are not constrained, are made to revolve accompanying output rotation of third ring gear R 3 , and thus revolve about third sun gear S 3  and fourth sun gear S 4  which are both fixed. Also, torque which constrains second sun gear S 2  acts through first connecting member M 1 , third sun gear  53 , third planet-pinions P 3 , and fourth sun gear S 4 . 
     (Third gear) In third gear, as shown in  FIG. 2 , second brake B 2  which was engaged in second gear is disengaged, and second clutch C 2  is engaged instead. Therefore, third gear is achieved through engagement of first clutch C 1  and second clutch C 2 . In third gear, reduced rotation from first planetary gearset G 1  is input into second ring gear R 2  of second planetary gearset G 2  by engagement of first clutch C 1 . Simultaneously, by engagement of second clutch C 2 , this reduced rotation is input into second sun gear S 2  of second planetary gearset G 2 . Thus, at second planetary gearset G 2 , by the same reduced rotation being input from second ring gear R 2  and second sun gear S 2 , second carrier PC 2  rotates integrally therewith, and a reduced rotation which is the same as the reduced rotation from planetary gearset G 1  is input to output gear  2  through second connecting member M 2 . The power flow in third gear is as shown in  FIG. 3C . Torque acts through first clutch C 1 , second clutch C 2 , and second connecting member M 2 , shown in bold lines, and first planetary gearset G 1  and second planetary gearset G 2 , shown in hatching. Third planetary gearset G 3  does not participate in transmission of torque. 
     (Fourth gear) In fourth gear, as shown in  FIG. 2 , second clutch C 2  which was engaged in third gear is disengaged, and third clutch C 3  is engaged instead. Therefore, fourth gear is achieved by engagement of first clutch C 1  and third clutch C 3 . In fourth gear, reduced rotation from first planetary gearset G 1  is input into second ring gear R 2  of second planetary gearset G 2  by engagement of first clutch C 1 . At the same time, input rotation from input shaft  1  is input into third carrier PC 3  of third planetary gearset G 3  through center member CM by engagement of third clutch C 3 . As a result, rotation of third sun gear S 3  is faster than output rotation of third ring gear R 3 , and this faster rotation of third sun gear S 3  is transmitted to second sun gear S 2  through first connecting member M 1 . 
     Thus, at second planetary gearset G 2 , reduced rotation from second ring gear R 2  is input, and faster rotation is input from second sun gear S 2 , and as a result, rotation which is a faster reduced rotation from second ring gear R 2  is output to output gear  2  from second carrier PC 2  through second connecting member M 2 . This faster reduced rotation from second ring gear R 2  is slower than the input rotation from input shaft  1 . The power flow in fourth gear is as shown in  FIG. 4A . Torque acts through first clutch C 1 , third clutch C 3 , center member CM, first member M 1 , and second member M 2 , shown in bold lines, and first planetary gearset G 1 , second planetary gearset G 2 , and third planetary gearset G 3  not including fourth sun gear S 4 , shown in hatching. 
     (Fifth gear) In fifth gear, as shown in  FIG. 2 , first clutch C 1  which was engaged in fourth gear is disengaged, and second clutch C 2  is engaged instead. Therefore, fifth gear is achieved by engagement of second clutch C 2  and third clutch C 3 . In fifth gear, reduced rotation from first planetary gearset G 1  is input into third sun gear S 3  through second sun gear S 2  and first connecting member M 1  by engagement of second clutch C 2 . Simultaneously, input rotation from input shaft  1  is input into third carrier PC 3  through center member CM by engagement of third clutch C 3 . 
     Thus, at third planetary gearset G 3 , input rotation is input into third carrier PC 3 , and reduced rotation is input into third sun gear S 3  from first planetary gearset G 1 , and as a result, rotation which is faster than the input rotation is output to output gear  2  from third ring gear R 3 . The power flow in fifth gear is as shown in  FIG. 4B . Torque acts through second clutch C 2 , third clutch C 3 , center member CM, and first connecting member M 1 , shown in bold lines, and first planetary gearset G 1 , second sun gear S 2 , and third planetary gearset G 3  not including fourth sun gear S 4 , shown in hatching. 
     (Sixth gear) In sixth gear, as shown in  FIG. 2 , second clutch C 2  which was engaged in fifth gear is disengaged, and second brake B 2  is engaged instead. Therefore, sixth gear is achieved by engagement of third clutch C 3  and second brake B 2 . In sixth gear, input rotation from input shaft  1  is input into third carrier PC 3  through center member CM of third planetary gearset G 3  by engagement of third clutch C 3 . Also, fourth sun gear S 4  of third planetary gearset G 3  is fixed to transmission case  3  by engagement of second brake B 2 . 
     Thus, at third planetary gearset G 3 , input rotation is input into third carrier PC 3 , and fourth sun gear S 4  is fixed to transmission case  3 , and as a result, rotation which is faster than the input rotation is output to output gear  2  from third ring gear R 3 . The power flow in sixth gear is as shown in  FIG. 4C . Torque acts through third clutch C 3 , second brake B 2 , and center member CM, shown in bold lines, and third planetary gearset G 3  not including third sun gear S 3 , shown in hatching. 
     (Reverse gear) As shown in  FIG. 2 , reverse gear, is achieved by engagement of second clutch C 2  and first brake B 1 . In reverse gear, reduced rotation from first planetary gearset G 1  is input into third sun gear S 3  through second sun gear S 2  and first connection member M 1  by engagement of second clutch C 2 . At the same time, by engagement of first brake B 1 , third carrier PC 3  is fixed to transmission case  3 . Thus, at third planetary gearset G 3 , normal-direction reduced rotation is input into third sun gear S 3 , and third carrier PC 3  is fixed to transmission case  3 , and as a result, reverse-direction rotation that has been reduced in speed is output to output gear  2  from third ring gear R 3 . 
     The power flow in reverse gear is as shown in  FIG. 5 . Torque acts through second clutch C 2 , first brake B 1 , first connecting member M 1 , and outer member OM, shown in bold lines, and first planetary gearset G 1 , second sun gear S 2 , and third planetary gearset G 3  except fourth sun gear S 4 , shown in hatching. 
     Below, in discussion of the automatic transmission according to the present invention based on  FIGS. 6 and 8 , the orientation of the automatic transmission in  FIGS. 6 and 8  is opposite to that of  FIGS. 1 ,  3  through  5 , and  7 . That is, an engine would be connected on the left side in  FIGS. 1 ,  3  through  5 , and  7  and power flows from left to right, whereas in  FIGS. 6 and 8 , an engine would be connected on the right side and therefore power flow would be from right to left. Referring to  FIG. 6 , input shaft  1  and a middle shaft  4  are disposed in transmission case  3  so that a rear end of input shaft  1  is supported in a front end of middle shaft  4  to form a fitting portion, such that input shaft  1  is coaxially rotatable relative to middle shaft  4 . Input shaft  1  and middle shaft  4  are supported to be individually and freely rotatable with respect to transmission case  3 . 
     A front end opening of transmission case  3  near input shaft  1  is closed by an oil pump which comprises a pump housing  5  and a pump cover  6 , and input shaft  1  is passed through the oil pump to be axially supported thereby. An end of input shaft  1  protruding from the oil pump is connected to an engine (not shown) via a torque converter (not shown). 
     A rear end of middle shaft  4  which is away from input shaft  1  is supported to be freely rotatable by a case end  7  at a rear end of transmission case  3 . A midway wall  8  is disposed approximately halfway axially inside transmission case  3 , and output gear  2  is supported thereon to be freely rotatable. A hollow shaft  9  is disposed in a center opening of midway wall  8 , and the fitting portion of input shaft  1  and middle shaft  4  is supported to be freely rotatable inside hollow shaft  9  by the center hole of midway wall  8 . 
     As shown in  FIGS. 6 and 7 , first planetary gearset G 1  is disposed in a space existing in a front portion of the automatic transmission between the oil pump comprised of pump housing  5  and pump cover  6  and midway wall  8 , and, third clutch C 3  is disposed so as to enclose first planetary gearset G 1 . Referring to first planetary gearset G 1 , first sun gear S 1  is serration fitted to a center boss portion  6   a  projecting from a rear of pump cover  6  to be permanently non-rotatable so as to function as a reaction force stopper, and first ring gear R 1  which is a rotation input member is joined to an outer perimeter of a flange  10  which extends radially outward away from the axis from input shaft  1 . 
     A clutch drum  11  extends radially outward away from the axis from a front end of middle shaft  4  near input shaft  1  and encloses first ring gear R 1 . Third clutch C 3 , which serves as a direct clutch, is disposed about an outer circumference of first planetary gearset G 1 , third clutch C 3  comprising a clutch pack  12  and a clutch piston  13  which will be discussed later. Clutch pack  12  comprises alternating clutch plates respectively splined to an inner circumference of clutch drum  11  and an outer circumference of first ring gear R 1 . First ring gear R 1  also serves as a clutch hub of third clutch C 3 . Further, clutch piston  13 , which is a clutch piston of third clutch C 3 , is disposed on a side of first planetary gearset G 1  away from the oil pump which is comprised of pump housing  5  and pump cover  6 , and clutch piston  13  is slidably fitted to a front end of middle shaft  4  and a cylinder  11   a  of clutch drum  11  which faces first planetary gearset G 1 . 
     Third clutch C 3  is engageable by third clutch piston  13  traveling to the right of  FIG. 6  after having received line pressure supplied via a fluid passage  14  from a control valve body. A connecting shell  53 , which is drum-shaped, extends radially outward away from the axis from a front end of hollow shaft  9 , continues on to enclose third clutch C 3 , and continues further so that a front end of connecting shell  53  is joined to first carrier PC 1 . First carrier PC 1 , as obvious from the previous explanation, constitutes a rotation output member of first planetary gearset G 1  which serves as a reduction planetary gearset. 
     As shown in  FIGS. 6 through 8 , first clutch C 1 , second clutch C 2 , first brake B 1 , and second brake B 2  are disposed in a space which exists between midway wall, 8  and case end  7 . 
     Second planetary gearset G 2  and third planetary gearset G 3  are disposed about middle shaft  4 , second planetary gearset G 2  being positioned nearer to input shaft  1  than third planetary gearset G 3 . Second sun gear S 2  of second planetary gearset G 2  and third sun gear S 3  of third planetary gearset G 3  are joined to form a single integral body by first connecting member M 1  and are supported to be freely rotatable by middle shaft  4 . A clutch drum  15  extends radially outward away from the axis from approximately halfway of hollow shaft  9 , continues on to extend axially toward the rear of transmission case  3 , and continues further somewhat past an outer circumference of second ring gear R 2 . First clutch C 1  is comprised of a clutch pack  16  and a clutch piston  19 . Clutch pack  16  is comprised of alternating clutch plates respectively splined to an inner circumference of clutch drum  15  and an outer circumference of second ring gear R 2 . 
     As discussed above, second clutch C 2  is disposed nearer to input shaft  1  than first clutch C 1  which is disposed on an outer circumference of second planetary gearset G 2 , so a clutch hub  17  which extends radially outward away from the axis is fixedly installed to an outer edge of an input shaft of second sun gear S 2 . Second clutch C 2  is comprised of a clutch pack  18  and a clutch piston  20  which will be discussed hereinafter. Clutch pack  18  is comprised of alternating clutch plates respectively splined to an inner circumference of clutch drum  15  and an outer circumference of clutch hub  17 . Further, clutch piston  19  of first clutch C 1  and clutch piston  20  of second clutch C 2  form a double piston which is disposed on a side of second clutch C 2  away from first clutch C 1 , clutch piston  20  being slidable on an inner side of clutch piston  19 . An outer circumference of clutch piston  19  is fitted to be freely slidable on an inner circumference of a cylinder  15   a  of clutch drum  15  which faces second planetary gearset G 2 . A plurality of fluid passages  21 , of which a representative fluid passage is shown in  FIG. 6 , are formed in midway wall  8  and hollow shaft  9 . First clutch C 1  and second clutch C 2  are individually engageable, by clutch piston  19  and clutch piston  20  traveling to the left of  FIG. 6  after receiving line pressure from individual fluid passages of the plurality of fluid passages  21 . 
     As mentioned before, third planetary gearset G 3  is a double-sun-gear planetary gearset, and third planet-pinions P 3  are relatively long in order to mesh with both third sun gear S 3  and fourth sun gear S 4 . However, a width of third ring gear R 3  does not need to be as large as that of third planet-pinions P 3 , and therefore third ring gear R 3  is fabricated so that a face width of the teeth thereof is smaller than a face width of the teeth of third planet-pinions P 3 . By third ring gear R 3  being meshed with third planet-pinions P 3  at an end portion of third planet-pinions P 3  near second planetary gearset G 2 , second connecting member M 2  which joins third ring gear R 3  and second carrier PC 2  of second planetary gearset G 2  can be designed smaller. An output drum  22  is disposed so as to enclose clutch drum  15  of first clutch C 1  and second clutch C 2 , and serves as an output member of the compound planetary gear train which comprises second planetary gearset G 2  and third planetary gearset G 3 . One end of output drum  22  is attached to an outer circumference of third ring gear R 3 , and another end thereof is attached to output gear  2 . 
     The output member of the compound planetary gear train is disposed outside the respective outer circumferences of first clutch C 1  and second clutch C 2  and inside the respective inner circumferences of first brake B 1  and second brake B 2 . Further, by disposing the output member of the compound planetary gear train in this way radially beyond the respective outer circumferences of first clutch C 1  and second clutch C 2  which are engageable to direct reduced rotation, as well as radially within the respective inner circumferences of first brake B 1  and second brake B 2 , both of which will be discussed in detail later, it is possible to form the output member, being output drum  22 , with a large diameter, which is favorable in terms of strength. Thus, according to the embodiment of the present invention, a thickness of output drum  22  is designed to be smaller than is generally so in the related art, while retaining sufficient strength characteristics. 
     As has been discussed, center member CM is disposed on third carrier PC 3  of third planetary gearset G 3  to extend radially inward toward the axis between third sun gear S 3  and fourth sun gear S 4 , and outer member OM is disposed on third carrier PC 3  at a position approximately halfway axially of third planet-pinions P 3  and extends radially outward away from the axis and along a rear face of third ring gear R 3 . Center member CM is connected to middle shaft  4 , and third carrier PC 3  is thereby connected to clutch drum  11  of third clutch C 3  via center member CM and middle shaft  4 . A brake hub  23  is joined to an outer circumference of outer member OM, and is disposed about an outer circumference of output drum  22  and extends toward the front of the transmission to within proximity of midway wall  8 . First brake B 1  is comprised of brake pack  24  and brake piston  25 . Brake pack  24  is comprised of alternating plates respectively splined to an inner circumference of brake hub  23  and an inner circumference of transmission case  3 . Brake piston  25  of first brake B 1  is fitted to the inside of transmission case  3  behind brake pack  24 , and first brake B 1  is appropriately engageable by brake piston  25 . 
     A brake hub  26  is disposed so as to overlap a rear end of brake hub  23 , and an end wall  26   a  of brake hub  26  extends inward toward the axis along and behind third planetary gearset G 3 , and an inner circumference of rear wall  26   a  of brake hub  26  is joined to fourth sun gear S 4  of third planetary gearset G 3 . Second brake B 2  is comprised of a brake pack  27  and a brake piston  28 . Brake pack  27  is comprised of alternating plates respectively splined to an inner circumference of transmission case  3  and an outer circumference of brake hub  26 . Brake piston  28  of second brake B 2  is fitted to the inside of transmission case  3  behind brake pack  27 , and second brake B 2  is appropriately engageable by brake piston  28 . 
     Therefore, first brake B 1  and second brake B 2  are respectively disposed around second clutch C 2  and around first clutch C 1  respectively, and first brake B 1  is disposed nearer to input shaft  1  and first planetary gearset G 1  than second brake B 2 . First brake B 1  and second brake B 2  are disposed axially in a row and nearer to second planetary gearset G 2  than third planetary gearset G 3 . 
     The disposition of second planetary gearset G 2 , third planetary gearset G 3 , first clutch C 1 , second clutch C 2 , first brake B 1 , and second brake B 2  in accordance with the embodiment of the present invention will now be discussed in more detail. As shown in  FIGS. 7 and 8 , first clutch C 1  and second clutch C 2 , which direct reduced rotation, are disposed radially beyond the compound planetary gear train which is comprised of second planetary gearset G 2  and third planetary gearset G 3 , that is, first clutch C 1  and second clutch C 2  are disposed outside and around the compound planetary gear train. First brake B 1  and second brake B 2  are respectively disposed radially beyond the circumferences of first clutch C 1  and second clutch C 2 , that is, first brake B 1  and second brake B 2  are individually disposed around second clutch C 2  and first clutch C 1  respectively, such that one of first clutch C 1  and second clutch C 2  and one of first brake B 1  and second brake B 2  overlap at least partially in the axial direction, and the other of first clutch C 1  and second clutch C 2  and the other of first brake B 1  and second brake B 2  overlap at least partially in the axial direction. 
     More specifically, second brake B 2  is disposed around the circumference of first clutch C 1  so that clutch pack  16  of first clutch C 1  and brake pack  27  of second brake B 2  greatly overlap in the axial direction. Also, second brake B 2  is disposed far enough toward the rear of the transmission, so that clutch pack  18  of second clutch C 2  and brake pack  24  of first brake B 1  overlap in the axial direction. Also, first clutch C 1  and second clutch C 2  are positioned in a row axially. 
     As shown in  FIG. 8 , working fluid to first brake B 1  is supplied through an opening member  51  which is disposed in the enclosing wall of transmission case  3 , and working fluid to second brake B 2  is supplied through an opening portion  52  which is disposed in case end  7 . At the same time, lubricating oil flows radially outward away from the axis from within middle shaft  4 , and is supplied to first clutch C 1 , second clutch C 2 , first brake B 1 , second brake B 2 , and other elements. With that in consideration, first clutch C 1 , second clutch C 2 , first brake B 1 , and second brake B 2  are positioned axially close together to facilitate the layout of fluid passages with respect to the axial direction. The fluid passage structure is thus simplified, especially with regard to lubricating oil. 
     Further, a one-way clutch OWC, which was omitted in the skeleton diagrams of  FIGS. 1 and 3  through  5 , is disposed between transmission case  3  and a front of brake hub  23  which constitutes first brake B 1 , as shown in  FIGS. 6 ,  7 , and  8 . Forward first gear is achievable even with first brake B 1  in a disengaged state due to one-way clutch OWC stopping rotation of third carrier PC 3  in one direction. However, while first gear can be achieved through provision of one-way clutch OWC, one-way clutch OWC allows reverse-direction rotation of third carrier PC 3  during engine braking. Engine braking is therefore not effective in this case. When engine braking is required, first brake B 1  is engaged to stop reverse-direction rotation of third carrier PC 3 . A counter shaft  29  is supported to be freely rotatable inside transmission case  3 , and is parallel to input shaft  1  and middle shaft  4 . A countergear  30  and a final drive pinion  31  are formed integrally with counter shaft  29 , countergear  30  meshing with output gear  2 , and final drive pinion  31  meshing with a differential gear assembly of drive wheels of a vehicle (not shown) are formed integrally with counter shaft  29 . 
     According to the embodiment of the present invention, referring to  FIGS. 6 through 8 , at a position which is radially beyond and along first clutch C 1  and second clutch C 2  which direct reduced rotation, one of first clutch C 1  and second clutch C 2  and one of first brake B 1  and second brake B 2  overlap at least partially in the axial direction, and the other of first clutch C 1  and second clutch C 2  and the other of first brake B 1  and second brake B 2  overlap at least partially in the axial direction. It therefore becomes possible to axially shorten the transmission case. And since it is possible to position first clutch C 1  and second brake B 2  axially closer to second clutch C 2  and first brake B 1  respectively, layout of fluid passages is facilitated in the axial direction and fluid passage structure is thereby simplified. 
     Further, the greater the overlap between one clutch and one brake (according to the embodiment, first clutch C 1  and second brake B 2 ) and between the other clutch and the other brake (according to the embodiment, second clutch C 2  and first brake B 1 ), the more noticeable the beneficial effects of the present invention are. It is therefore preferable to dispose first clutch C 1  and second brake B 2  as well as second clutch C 2  and first brake B 1  to overlap as much as can be allowed by the particular arrangement of a given transmission assembly. 
     With respect to the radial dimension of the transmission assembly, second planetary gearset G 2  and third planetary gearset G 3  which constitute the compound planetary gear train are single-pinion, so the compound planetary gear train can be designed with a smaller diameter. Moreover, by making third planetary gearset G 3  of the compound planetary gearset a double-sun-gear planetary gearset having two sun gears, it is possible for second ring gear R 2  to serve as the input member of the compound planetary gear train for reduced rotation from the reduction planetary gearset, being first planetary gearset G 1 . Compared to a sun gear serving as an input member, there is less tangential stress present with a ring gear acting as the input member, and is therefore advantageous with respect to a number of points including gear strength, gear life, and carrier rigidity, and it is possible to make a diameter of the compound planetary gear train smaller. 
     Also, in an automatic transmission according to the embodiment of the present invention, output drum  22  acts as an output member of the compound planetary gear train which comprises single-pinion planetary gearset G 2  and double-sun-gear planetary gearset G 3 , and is disposed radially beyond the respective outer circumferences of first clutch C 1  and second clutch C 2  as well as radially within the respective inner circumferences of first brake B 1  and second brake B 2 . This allows output drum  22  to be made with a larger diameter, which is an advantage with respect to strength. With a larger diameter, output drum  22  can then be designed with a smaller thickness and still retain sufficient strength for transmitting high torque. This makes it possible to more effectively design a smaller transmission. 
     This application is based on a prior Japanese Patent Application No. P2002-207345. The entire contents of a Japanese Patent Application No. P2002-207345 with a filing date of Jul. 16, 2002 are hereby incorporated by reference. 
     Although the invention has been described above by reference to an embodiment of the invention, the invention is not limited to the embodiment described above. Modifications and variations of the embodiment described above will occur to those skilled in the art in light of the above teachings. 
     For example, the present invention can be applied in an instance where first planetary gearset G 1  is a double-pinion planetary gearset with two sets of planet-pinion gears where the rotation input member is first carrier PC 1 , and the rotation output member is first ring gear R 1 . 
     Also, the present invention can be applied to an automatic transmission where there are more than three planetary gearsets. 
     The scope of the invention is defined with reference to the following claims.