Abstract:
A first planetary gear unit is arranged to convert an input rotation from an input shaft to an output rotation whose speed is lower than that of the input rotation. A first unit including second and third planetary gear units is arranged between the first planetary gear unit and the output shaft. The first unit manages the first planetary gear unit&#39;s output rotation and provides the output shaft with seven types of rotation which correspond to 6-forward speed and one reverse positions. A second unit is arranged between the first planetary gear unit and the first unit to manage a power transmission therebetween. One of the second and third planetary gear units is of a double ring type, each being meshed with the sun gear and inside and outside ring gears, and a pinion carrier carrying the pinions. The pinion carrier is connected to the output shaft to rotate therewith.

Description:
BACKGROUND OF THE INVENTION 
   1. Field of the Invention 
   The present invention relates in general to automatic transmissions for wheeled motor vehicles, and more particularly to a speed change mechanism of the automatic transmissions, which has 6-forward speed and one reverse positions. 
   2. Field of the Related Art 
   One of the speed change mechanisms of the above-mentioned type is shown in FIG. 5 of Japanese Laid-open Patent Application (Tokkaihei) 4-219553. The speed change mechanism of this published application comprises generally an input shaft, a planetary gear train including a single pinion type gear unit and two single pinion type gear units, three clutches, two brakes and an output shaft. For achieving 6-forward speed and one reverse portions, the three clutches and two brakes are selectively engaged and disengaged in a given manner. 
   To provide the speed change mechanism with an overdrive speed position, it is necessary to apply an input force to a pinion carrier and a ring gear of the planetary gear train. However, if this necessity is made in an arrangement wherein the input and output shaft are arranged coaxially, the single pinion type planetary gear, which has only three rotational members, can not provide an input power path for the pinion carrier and the ring gear. For providing such power path, the input and output shafts have to be arranged on different axes which are parallel with each other. However, this arrangement tends to induce a bulky construction of an associated automatic transmission, particularly, the size in a radial direction. 
   In order to solve such drawback in size, the same published application shows in  FIG. 3  another speed change mechanism. In this mechanism, the input and output shafts are arranged coaxially, and in place of the above-mentioned planetary gear train including the two single pinion type gear units, a ravigneawx type complex planetary gear train is used. This gear train has an arrangement wherein two planetary gear units are arranged having their double pinions engaged with respective sun gears. 
   In this ravigneawx type transmission, one of the planetary gear units has a double pinion type. This means increase in number of portions where gear meshing is made for achieving power transmission. However, increase in number of such portions tends to induce undesirable gear noise and vibration. For preventing such noise and vibration, highly accurate and thus expensive machining and assembling process is needed. 
   In view of the above-mentioned drawbacks, Japanese Laid-open Patent Application 2001-349390 has proposed in FIGS. 9, 20, 13 and 34, a 6-forward speed and one reverse speed change mechanism which comprises coaxially arranged input and output shafts, a planetary gear train including one speed reduction planetary gear unit and two single pinion planetary gear units, three clutches and two brakes. 
   SUMMARY OF THE INVENTION 
   In the speed change mechanism of the application 2001-349390, even if respective planetary gear ratios (viz., number of teeth of sun gear/number of teeth of ring gear) of the speed reduction planetary gear unit and the two single pinion planetary gear unit are set in a preferable range, it tends to occur that a ratio coverage from 1 st  forward speed to 6 th  forward speed (viz., gear ratio at 1 st  forward speed/gear ratio at 6 th  forward speed) becomes narrowed and the rate between gear ratio at reverse position and gear ratio at 1 st  forward speed (viz., gear ratio at reverse position/gear ratio at 1 st  forward speed: which will be referred to “1-R ratio” for ease of description) is not set to a suitable value. 
   As is known, narrowing the ratio coverage debases the essential feature of 6-forward speed and one reverse position transmission, deteriorating freedom in selecting speeds. Furthermore, if the ratio coverage fails to have a satisfied maximum value, fuel consumption and drivability of an associated motor vehicle become poor. 
   Furthermore, if the “1-R ratio” fails to be set at a desired value, that is, for example, if the “1-R ratio” is set at a smaller value, drivability of the vehicle becomes poor because respective output torques produced at 1 st  forward speed and reverse position are largely different. Furthermore, if the gear ratio for 1 st  forward speed is set at a desired value, the gear ratio for reverse position is inevitably set to have a higher speed side. In this case, it is necessary to increase the accelerator open degree largely for obtaining a satisfied torque in the reverse position. On the contrary, if the gear ratio for reverse position is set at a desired value, the gear ratio for 1 st  forward speed is inevitably set to have a very low speed side. In this case, fuel consumption and drivability of the vehicle become poor. 
   In view of the above-mentioned various facts in conventional technique, the present invention aims to provide a speed change mechanism of an automatic transmission, in which input and output shafts are arranged coaxially for achieving a compact construction of the mechanism, a freedom in selecting gear ratio is sufficiently obtained and drivability and fuel consumption of an associated motor vehicle are improved. 
   According to a first aspect of the present invention, there is provided a speed change mechanism of an automatic transmission for achieving 6-forward speed and one reverse positions, which comprises an input shaft adapted to be driven by a power source; an output shaft coaxially arranged with the input shaft; a first planetary gear unit which converts an input rotation from the input shaft to an output rotation whose speed is lower than that of the input rotation; and a first unit including second and third planetary gear units and arranged between the first planetary gear unit and the output shaft, the first unit being arranged to provide, by managing the output rotation from the first planetary gear unit, the output shaft with seven types of rotation which respectively correspond to the 6-forward speed and one reverse positions; and a second unit arranged between the first planetary gear unit and the first unit to manage a power transmission therebetween, wherein one of the second and third planetary gear units is of a double ring type which comprises a sun gear powered by the first planetary gear unit, inside and outside ring gears concentrically disposed around the sun gear, pinions each being meshed with the sun gear and the inside and outside ring gears, and a pinion carrier carrying the pinions, the pinion carrier being connected to the output shaft to rotate therewith. 
   According to a second aspect of the present invention, there is provided a speed change mechanism of an automatic transmission for achieving 6-forward speed and one reverse positions, which comprises an input shaft adapted to be driven by a power source; an output shaft coaxially arranged with the input shaft; a first planetary gear unit including a first sun gear fixed to a case of the transmission, a first ring gear connected to the input shaft, first pinions each being meshed with both the first sun gear and the first ring gear and a first pinion carrier carrying the first pinions; a second planetary gear unit including a second sun gear, a second ring gear, second pinions each being meshed with both the second sun gear and the second ring gear and a second pinion carrier carrying the second pinions; a third planetary gear unit including a third sun gear, third and fourth ring gears, third pinions each being meshed with the third sun gear and the third and fourth ring gears and a third pinion carrier carrying the third pinions; a first connecting member connecting the second pinion carrier to the fourth ring gear; a second connecting member connecting the first ring gear to the third ring gear; a third clutch incorporated with the second connecting member; a third connecting member connecting the second ring gear to the output shaft; a fourth connecting member connecting the first pinion carrier to the second sun gear; a second clutch incorporated with the fourth connecting member; a fifth connecting member connecting the first pinion carrier to the third sun gear; a first clutch incorporated with the fifth connecting member; a center member connecting the third pinion carrier to the third connecting member, the center member extending radially outward from the third pinion carrier to the third connecting member through a space defined between the third and fourth ring gears; a first brake which is able to brake the second pinion carrier; and a second brake which is able to brake the second sun gear. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
     Other objects and advantages of the present invention will become apparent from the following description when taken in conjunction with the accompanying drawings, in which: 
       FIG. 1  is a schematic view of a speed change mechanism of an automatic transmission, which is an embodiment of the present invention; 
       FIG. 2  is a table showing ON/OFF condition of each friction element with respect to six forward speed and reverse positions established in the speed change mechanism of the present invention; 
       FIG. 3  is an alignment chart representing the dynamic feature of the speed change mechanism of the invention; 
       FIGS. 4  to  9  are views similar to  FIG. 1 , but showing torque transmitting paths established in 1 st , 2 nd , 3 rd , 4 th , 5 th  and 6 th  forward speeds of the speed change mechanism of the present invention, respectively; 
       FIG. 10  is a view similar to  FIG. 1 , but showing a torque transmitting path established in a reverse position of the speed change mechanism of the present invention; 
       FIG. 11  a view similar to  FIG. 10 , but showing a speed change mechanism equipped with a ravigneawx type planetary gear train; 
       FIG. 12  is a comparative table showing the performance of the speed change mechanism of the invention and that of the speed change mechanism equipped with the ravigneawx type planetary gear train; 
       FIG. 13  is a comparative table showing respectively the ratio coverage and 1-R ratio in case of the speed change mechanism of the invention, a speed change mechanism including a single pinion type planetary gear unit and a ravigneawx type planetary gear unit and a speed change mechanism including a single speed reduction planetary gear unit and two single pinion type planetary gear units; 
       FIG. 14  is a comparative graph showing a change of the ratio coverage in case of the speed change mechanism of the invention, the speed change mechanism including the single pinion type planetary gear unit and the ravigneawx type planetary gear unit and the speed change mechanism including the single speed reduction planetary gear unit and the two single pinion type planetary gear units; and 
       FIG. 15  is a graph showing a change of the 1-R ratio in case of the speed change mechanism of the invention, the speed change mechanism including the single pinion type planetary gear unit and the ravigneawx type planetary gear unit and the speed change mechanism including the single speed reduction planetary gear unit and the two single pinion type planetary gear units. 
   

   DETAILED DESCRIPTION OF THE INVENTION 
   In the following, an embodiment of the present invention will be described in detail with reference to the accompanying drawings. 
     FIG. 1  shows schematically a speed change mechanism of an automatic transmission, which is the embodiment of the present invention. 
   In the drawing, denoted by “Input” is an input shaft into which a torque is applied from an engine (not shown) through a torque converter (not shown), and denoted by “Output” is an output shaft from which a torque is led to driven road wheels (not shown) of an associated motor vehicle through a final gear (not shown). These input and output shafts are arranged coaxially, as shown. 
   In the side of the input shaft “Input”, there is arranged a first planetary gear unit G 1  which converts an input rotation from the input shaft “Input” to an output rotation whose speed is lower than that of the input rotation. While, in the side of the output shaft “Output”, there are arranged a second planetary gear unit G 2  and a double ring type third planetary gear unit G 3 . 
   Denoted by “C 1 ” is a first clutch, “C 2 ” is a second clutch, “C 3 ” is a third clutch, “B 1 ” is a first brake, “B 2 ” is a second brake, “M 1 ” is a first connecting member, “M 2 ” is a second connecting member, “M 3 ” is a third connecting member, “M 4 ” is a fourth connecting member and “M 5 ” is a fifth connecting member. 
   The first planetary gear unit G 1  is of a single pinion type and comprises a first sun gear S 1 , a first ring gear R 1 , first pinions P 1  each meshed with both the first sun gear S 1  and the first ring gear R 1  and a first pinion carrier PC 1  carrying the first pinions P 1 . 
   The second planetary gear unit G 2  is of a single pinion type and comprises a second sun gear S 2 , a second ring gear R 2 , second pinions P 2  each meshed with both the second sun gear S 2  and the second ring gear R 2  and a second pinion carrier PC 2  carrying the second pinions P 2 . 
   The third planetary gear unit G 3  is of a double ring gear type and comprises a third sun gear S 3 , two, viz., third and fourth ring gears R 3  and R 4 , third pinions P 3  each meshed with the third sun gear S 3  and the third and fourth ring gears R 3  and R 4 , and a third pinion carrier PC 3  which carries the third pinions P 3  and has a center member CM arranged between third and fourth ring gears R 3  and R 4 . The center member CM is connected to the third pinion carrier PC 3  at spaced positions of the third pinions P 3  which are placed in the vicinity of a periphery of the third pinion carrier PC 3 . 
   The first clutch C 1  functions to selectively engage and disengage the first pinion carrier PC 1  of the first planetary gear unit G 1  to and from the third sun gear S 3  of the third planetary gear unit G 3 . The second clutch C 2  functions to selectively engage and disengage the first pinion carrier PC 1  of the first planetary gear unit G 1  to and from the second sun gear S 2  of the second planetary gear unit G 2 . The third clutch C 3  functions to selectively engage and disengage the first ring gear R 1  of the first planetary gear unit G 1  to and from the third ring gear R 3  of the third planetary gear unit G 3 . 
   The first brake B 1  functions to selectively brake the second pinion carrier PC 2  of the second planetary gear unit G 2 . The second brake B 2  functions to selectively brake the second sun gear S 2  of the second planetary gear unit G 2 . 
   The first connecting member M 1  connects the second pinion carrier PC 2  to the fourth ring gear of the third planetary gear unit G 3 . The second pinion carrier PC 2  is connected to the first brake B 1 . 
   The second connecting member M 2  connects the third ring gear R 3  of the third planetary gear unit G 3  to the first ring gear R 1  of the first planetary gear unit G 1 . The third clutch C 3  is incorporated with the second connecting member M 2 . 
   The third connecting member M 3  connects the second ring gear R 2  to the output shaft “Output”. 
   The center member CM connects the third pinion carrier PC 3  to the third connecting member M 3  which radially outwardly extending in a space defined between the third and fourth ring gears R 3  and R 4 , as shown. 
   The fourth connecting member M 4  connects the second sun gear S 2  of the second planetary gear unit G 2  to the first pinion carrier PC 1  of the first planetary gear unit G 1 . The second clutch C 2  is incorporated with the fourth connecting member M 4 . The fourth connecting member M 4  is connected to the second brake B 2 , as shown. 
   The fifth connecting member M 5  connects the third sun gear S 3  of the third planetary gear unit G 3  to the fourth connecting member M 4 . The first clutch C 1  is incorporated with the fifth connecting member M 5 . 
   To the first to third clutches C 1 , C 2  and C 3  and first and second brakes B 1  and B 2 , there are respectively connected hydraulically actuating devices (not shown) of a speed change control device. 
   When the above-mentioned clutches C 1 , C 2  and C 3  and brakes B 1  and B 2  are engaged or disengaged by the hydraulically actuating devices in a manner as is indicated by the table of  FIG. 2 , 6-forward speed and one reverse positions are selectively obtained. As the speed change control device, a hydraulic type, electronic type or a combination of these types may be used. 
   It is to be noted that the table of  FIG. 2  depicts the engaged/disengaged condition of each frictional element with respect to a speed position (or reverse position) assumed by the speed change mechanism. Engaged condition is represented by a black circle, and disengaged condition is represented by a blank. 
   In the following, operation of the speed change mechanism will be described with reference to the alignment chart of FIG.  3  and the torque transmitting paths of  FIGS. 4  to  10  of the drawings. In  FIGS. 4  to  10 , established torque transmitting path is indicated by hatched blocks and thicker lines. 
   1 st  Forward Speed 
   As is seen from the table of  FIG. 2 , in 1 st  forward speed of the speed change mechanism, both the first clutch C 1  and first brake B 1  are engaged respectively. 
   As is seen from  FIG. 4 , in this 1 st  forward speed, due to the engagement of the first clutch C 1 , a reduced speed rotation in normal direction from the first planetary gear unit G 1  is inputted to the third sun gear S 3  of the third planetary gear unit G 3  through the fifth connecting member M 5 . 
   In the second planetary gear unit G 2 , due to engagement of the first brake B 1 , the second pinion carrier PC 2  is fixed to the case. In addition, the fourth ring gear R 4  of the third planetary gear unit G 3 , that is connected to the second pinion carrier PC 2  through the first connecting member M 1 , is also fixed to the case. 
   Thus, although the third sun gear S 3  of the third planetary gear unit G 3  is applied with the reduced speed rotation in the normal direction, the fixed condition of the fourth ring gear R 4  induces that a further reduced speed rotation is outputted from the third pinion carrier PC 3  to the output shaft “Output” through the center member CM. 
   Accordingly, in the 1 st  forward speed, as is shown in the alignment chart of  FIG. 3 , there is established a characteristic line “1st” which connects an engaging point of the first clutch C 1  through which the reduced speed rotation from the first planetary gear unit G 1  is transmitted to the third sun gear S 3  of the third planetary gear unit G 3  and an engaging point of the first brake B 1  by which a rotation of the second pinion carrier PC 2  of the second planetary gear unit G 2  is stopped. That is, rotation inputted from the input shaft “Input” is outputted to the output shaft “Output” while being reduced in speed. 
   2 nd  Forward Speed 
   As is seen from the table of  FIG. 2 , in 2 nd  forward speed, both the first clutch C 1  and the second brake B 2  are engaged respectively. 
   As is seen from  FIG. 5 , in this 2 nd  forward speed, due to engagement of the first clutch C 1 , a reduced speed rotation in normal direction from the first planetary gear unit G 1  is inputted to the third sun gear S 3  of the third planetary gear unit G 3  through the fifth connecting member M 5 , and a rotation is transmitted to the output shaft “Output” from the third pinion carrier PC 3  through the center member CM. 
   In the second planetary gear unit G 2 , due to engagement of the second brake B 2 , the second sun gear S 2  is fixed to the case. In addition, into the second pinion P 2 , there is inputted a reduced speed rotation in normal direction from the fourth ring gear R 4  of the third planetary gear unit G 3  through the first connecting member M 1 . Due to the fixed condition of the second sun gear S 2 , the second ring gear R 2  is force to rotate in normal direction at an increased speed. The rotation of the second ring gear R 2  is transmitted to the third connecting member M 3 . 
   Accordingly, in the third planetary gear unit G 3 , a reduced speed rotation in normal direction is transmitted from the third pinion carrier PC 3  to the unit G 3  through the center member CM, and in the second planetary gear unit G 2 , an increased speed rotation in normal direction is transmitted from the second ring gear R 2  to the unit G 2  through the third connecting member M 3 , so that a rotation that is higher than that in the 1 st  forward speed is transmitted to the output shaft “Output”. 
   Accordingly, in 2 nd  forward speed, as is shown in the alignment chart of  FIG. 3 , there is established a characteristic line “2nd” which connects an engaging point of the first clutch C 1  through which the reduced speed rotation from the first planetary gear unit G 1  is transmitted to the third sun gear S 3  of the third planetary gear unit G 3  and an engaging point of the second brake B 2  by which a rotation of the second sun gear S 2  of the second planetary gear unit G 2  is stopped. That is, rotation inputted from the input shaft “Input” is outputted to the output shaft “Output” while being increased in speed to a value higher than that in 1 st  forward speed. 
   3 rd  Forward Speed 
   As is seen from the table of  FIG. 2 , in 3 rd  forward speed, both the first and second clutches C 1  and C 2  are engaged respectively. 
   As is seen from  FIG. 6 , in this 3 rd  forward speed, due to engagement of the first clutch C 1 , a reduced speed rotation in normal direction from the first planetary gear unit G 1  is inputted to the third sun gear S 3  of the third planetary gear unit G 3  through the fifth connecting member M 5 , and a rotation is transmitted to the output shaft “Output” from the third pinion carrier PC 3  through the center member CM. At the same time, due to engagement of the second clutch C 2 , the reduced speed rotation in normal direction from the first planetary gear unit G 1  is transmitted to the second sun gear S 2  of the second planetary gear unit G 2  through the fourth connecting member M 4 , and a rotation is transmitted to the output shaft “Output” from the second ring gear R 2  through the third connecting member M 3 . 
   Accordingly, in 3 rd  forward speed, as is seen from the alignment chart of  FIG. 3 , there is established a characteristic line “3rd” which connects an engaging point of the first clutch C 1  through which the reduced speed rotation from the first planetary gear unit G 1  is transmitted to the third sun gear S 3  of the third planetary gear unit G 3  and an engaging point of the second clutch C 2  through which the reduced speed rotation from the first planetary gear unit G 1  is transmitted to the second sun gear S 2  of the second planetary gear unit G 2 . That is, rotation inputted from the input shaft “Input” is outputted to the output shaft “Output” while being increased in speed to a value higher than that in 2 nd  forward speed. 
   4 th  Forward Speed 
   As is seen from the table of  FIG. 2 , in 4 th  forward speed, both the first clutch C 1  and the third clutch C 3  are engaged respectively. 
   As is seen from  FIG. 7 , in 4 th  forward speed, due to engagement of the first clutch C 1 , a reduced speed rotation in normal direction from the first planetary gear unit G 1  is inputted to the third sun gear S 3  of the third planetary gear unit G 3 , and, due to engagement of the third clutch C 3 , a rotation of the input shaft “Input” is inputted to the third ring gear R 3  of the third planetary gear unit G 3  through the second connecting member M 2 . 
   Accordingly, in 4 th  forward speed, the third sun gear S 3  is applied with a reduced speed rotation and the third ring gear R 3  is applied with an increased speed rotation, and thus, a rotation provided by increasing the reduced speed rotation of the third sun gear S 3 , that is lower than that of the input rotation), is outputted to the output shaft “Output” from the third pinion carrier PC 3  through the center member CM. 
   Accordingly, in 4 th  forward speed, as is seen from the alignment chart of  FIG. 3 , there is established a characteristic line “4th” which connects an engaging point of the first clutch C 1  through which the reduced speed rotation from the first planetary gear unit G 1  is transmitted to the third sun gear S 3  of the third planetary gear unit G 3  and an engaging point of the third clutch C 3  through which the rotation of the input shaft “Input” is inputted to the third ring gear R 3  of the third planetary gear unit G 3 . That is, rotation inputted from the input shaft “Input” is outputted to the output shaft “Output” while being increased in speed to a value higher than that in 3 rd  forward speed. 
   5 th  Forward Speed 
   As is seen from the table of  FIG. 2 , in 5 th  forward speed, both the second clutch C 2  and the third clutch C 3  are engaged respectively. 
   As is seen from  FIG. 8 , in 5 th  forward speed, due to engagement of the second clutch C 2 , the reduced speed rotation from the second planetary gear unit G 2  is inputted to the second sun gear S 2  of the second planetary gear unit G 2  through the fourth connecting member M 4 . and at the same time, due to engagement of the third clutch C 3 , the rotation of the input shaft “Input” is inputted to the third ring gear R 3  of the third planetary gear unit G 3  through the second connecting member M 2 . 
   Accordingly, in 5 th  forward speed, the reduced speed rotation is applied to the second sun gear S 2  and the rotation of the input shaft “Input” is inputted to the second pinion carrier PC 2  through the first connecting member M 1 , and thus, a rotation that is higher in speed than the input shaft “Input” is outputted from the second ring gear R 2  to the output shaft “Output” through the third connecting member M 3 . 
   Accordingly, in 5 th  forward speed, as is seen from the alignment chart of  FIG. 3 , there is established a characteristic line “5th” which connects an engaging point of the second clutch C 2  through which the reduced speed rotation from the first planetary gear unit G 1  is inputted to the second sun gear S 2  of the second planetary gear unit G 2  and an engaging point of the third clutch C 3  through which the rotation of the input shaft “Input” is inputted to the third ring gear “R 3 ” of the third planetary gear unit G 3 . That is, rotation inputted from the input shaft “Input” is outputted to the output shaft “Output” while being increased in speed to a value higher than that in 4 th  forward speed. 
   6 th  Forward Speed 
   As is seen from the table of  FIG. 2 , in 6 th  forward speed, both the third clutch C 3  and the second brake B 2  are engaged respectively. 
   As is seen from  FIG. 9 , in 6 th  forward speed, due to engagement of the third clutch C 3 , the rotation of the input shaft “Input” is inputted to the third ring gear R 3  of the third planetary gear unit G 3  through the second connecting member M 2 , and at the same time, due to engagement of the second brake B 2 , the second sun gear S 2  of the second planetary gear unit G 2  is fixed to the case. 
   Accordingly, in 6 th  forward speed, the rotation of the input shaft “Input” is inputted to the second pinion carrier PC 2  from the fourth ring gear R 4  of the third planetary gear unit G 3  through the first connecting member M 1 , and due to fixing of the second sun gear S 2 , a rotation that is much higher in speed than the input shaft “Input” is outputted from the second ring gear R 2  to the output shaft “Output” through the third connecting member M 3 . 
   Accordingly, in 6 th  forward speed, as is seen from the alignment chart of  FIG. 3 , there is established a characteristic line “6th” which connects an engaging point of the third clutch C 3  through the rotation of the input shaft “Input” is inputted to the third ring gear R 3  of the third planetary gear unit G 3  and an engaging point of the second brake B 2  through which the second sun gear S 2  of the second planetary gear unit G 2  is fixed to the case. That is, rotation inputted from the input shaft “Input” is outputted to the output shaft “Output” while being increased in speed to a value higher than that in 5 th  speed. 
   Reverse position 
   As is seen from the table of  FIG. 2 , in reverse position, both the second clutch C 2  and the first brake B 1  are engaged respectively. 
   As is seen from  FIG. 10 , in reverse position, due to engagement of the second clutch C 2 , the reduced speed rotation from the first planetary gear unit G 1  is inputted to the second sun gear S 2  of the second planetary gear unit G 2 , and due to fixing of the first brake B 1 , the second pinion carrier P 2  of the second planetary gear unit G 2  is fixed to the case. 
   Accordingly, in reverse position, the second sun gear S 2  is applied with a reduced speed rotation in reversed direction, due to fixing of the second pinion carrier PC 2  to the case, a reduced speed rotation in reversed direction is inputted to the output shaft “Output” from the second ring gear R 2  through the third connecting member M 3 . 
   Accordingly, in reverse position, as is seen from the alignment chart of  FIG. 3 , there is established a characteristic line “Rev” which connects an engaging point of the second clutch C 2  through which the reduced speed rotation from the first planetary gear unit G 1  is inputted to the second sun gear S 2  of the second planetary gear unit G 2  and an engaging point of the first brake B 1  through which the rotation of the second pinion carrier PC 2  of the second planetary gear unit G 2  is stopped. That is, rotation inputted from the input shaft “Input” is outputted to the output shaft “Output” while being decreased in speed and reversed in rotation direction. 
   In the following, advantageous features of the speed change mechanism of the present invention will be described with the aid of the accompanying drawings. 
   For clarifying the features of the invention, two conventional speed change mechanisms will be also briefly described, which are the mechanism shown in Japanese Laid-open Patent Application (Tokkaihei) 4-219553 and the mechanism shown in Japanese Laid-open Patent Application 2001-349390. That is, the mechanism of 4-219553 is a speed change mechanism equipped with a ravigneawx type planetary gear train, the mechanism of 2001-349390 is a speed change mechanism comprising a planetary gear train including one speed reduction planetary gear unit and two single pinion planetary gear units. 
     FIG. 11  is a schematic view of the speed change mechanism of 4-219553. As shown, the mechanism comprises a single pinion planetary gear unit G 4  and a ravigneawx type planetary gear train G 5 . The gear unit G 4  comprises a sun gear S 4 , a ring gear R 4 , pinions P 4  meshed with both the sun gear S 4  and the ring gear R 4 , and a pinion carrier PC 4  carrying the pinions P 4 . The ravigneawx type planetary gear train G 5  comprises two sun gears S 5   a  and S 5   b , two groups of pinions P 5   a  and P 5   b  respectively meshed with the sun gears S 5   a  and S 5   b , a ring gear R 5  and a pinion carrier PC 5 . 
   (1) Compactness 
   In order to provide the speed change mechanism with an over-drive speed, application of input to the pinion carrier and ring gear is usually needed. In the mechanism of ravigneawx type of  FIG. 11 , the input and output shafts “Input” and “Output” are arranged coaxially establishing the needed input application to the pinion carrier PC 5  and ring gear R 5 . Thus, this mechanism establishes the over-drive speed without increasing the radial size thereof. 
   Also in the speed change mechanism of the present invention, the over-drive speed is established without increasing the radial size thereof for the same reason as has just mentioned hereinabove. That is, as is seen in  FIG. 1 , the combined planetary gear train including the single pinion type second planetary gear unit G 2  and the double ring gear type third planetary gear unit G 3  has five connecting members, which are the first connecting member M 1  extending between the fourth ring gear R 4  and second pinion carrier PC 2 , the second connecting member M 2  extending between the third ring gear R 3  and third clutch C 3 , the third connecting member M 3  extending between the second ring gear R 2  and the output shaft “Output”, the fourth connecting member M 4  extending between the second sun gear S 2  and second clutch C 2  and the fifth connecting member M 5  extending between the third sun gear S 3  and the first clutch C 1 . Due to provision of the center member CM which extends radially outward between the third ring gear R 3  and fourth ring gear R 4 , torque output is easily achieved. Thus, in the speed change mechanism of the invention, higher speeds that need the force input to the pinion carrier and ring gear are easily obtained with a compact size. 
   (2) Noise Reduction 
   In the ravigneawx type speed change mechanism, two groups of pinions P 5   a  and P 5   b  are provided in the gear train G 5  as is seen from FIG.  11 . However, employment of the two groups of pinions tends to produce noises and vibrations due to increase in number of contact points where gear meshing is made. 
   In the speed change mechanism of the present invention, there is no means that correspond to the two groups of pinions P 5   a  and P 5   b . Thus, undesired noises and vibrations caused by such groups are not produced. 
   (3) Widening of Gear Ratio Coverage 
     FIG. 12  is a table showing the performance of the speed change mechanism of the present invention and that of the speed change mechanism equipped with the ravigneawx type planetary gear train. INVENTION-I and INVENTION-II are speed change mechanisms of the present invention, which have different planetary gear ratio, and RAVIGNEAWX-I and RAVIGNEAWX-II are the speed change mechanism of the ravigneawx type, which have different planetary gear ratio, as is shown in the table. In INVENTION-I and INVENTION-II, α 1 , α 2  and α 3  denote planetary gear ratios of the first, second and third planetary gear units G 1 , G 2  and G 3  respectively, and in RAVIGNEAWX-I and RAVIGNEAWX-II, α 1  denotes a planetary gear ratio of the planetary gear unit G 4 , α 2  denotes a planetary gear ratio of the side of the sun gear S 5   a  of the planetary gear unit G 5  and α 3  denotes a planetary gear ratio of the side of the other sun gear S 5   b  of the gear unit G 5 . 
   First, the speed change mechanism of ravigneawx type will be considered. In general, in this speed change mechanism, the radial size of the part of a case where the planetary gear unit G 5  is installed increases when the planetary gear ratio “α” (viz. number of teeth of sun gear/number of teeth of ring gear) is within a commonly employed range from 0.35 to 0.65 and an appropriate arrangement is employed wherein the gear spacing ratio lowers as the gear change has a higher speed position. That is, in the ravigneawx type, when the planetary gear ratio “α 1 ” is set at a small value, it becomes impossible to widen the gear ratio for achieving the above-mentioned conditions. While, when the planetary gear ratio “α 3 ” is set at a small value (viz., 0.37 in case of RAVIGNEAWX-I and 0.35 in case of RAVIGNEAWX-II), widening of the gear ratio is achieved. However, as is seen from  FIG. 11 , due to provision of the connecting members around the axis of the ravigneawx type planetary gear train G 5 , there is a limit in reducing the diameter of the sun gear S 5   b . Thus, for widening the gear ratio, that is, for causing the planetary gear ratio “α 3 ” to have a small value, it is necessary to increase the diameter of the ring gear R 5 . However, increase in diameter of the ring gear R 5  brings about the radial enlargement of the case in which the gear train G 5  is installed. This undesirable phenomenon becomes much severe when the transmission to which the speed change mechanism of the ravigneawx type is applied is of a type for use with a front engine rear drive motor vehicle. In short, in case of the ravigneawx type, it is difficult to widen the gear ratio without increasing the size of the transmission case. 
   While, in the speed change mechanism of the present invention, the above-mentioned conditions are satisfied even when the planetary gear ratio “α 3 ” is set at a relatively large value (for example, 0.59 in INVENTION-1). Accordingly, as is seen from  FIG. 1 , even if the connecting members are arranged around the axis of the third planetary gear unit G 3 , it is unnecessary to reduce the diameter of the third sun gear S 3 . That is, it is unnecessary to increase the diameter of the third and fourth ring gears R 3  and R 4 , and thus, the undesired radial expansion of the transmission case is not induced. 
   (4-1) Ratio coverage without considering the gear spacing ratio 
   In general, the range of planetary gear ratio “α” is from 0.35 to 0.65. Preferably, the range is from 0.38 to 0.60. 
     FIG. 13  is a comparative table showing at the upper row respective ratio coverage values (viz., gear ratio at 1 st  forward speed/gear ratio at 6 th  forward speed) of the speed change mechanism of the invention, the ravigneawx type speed change mechanism of Japanese Laid-open Patent Application 4-219553 and the speed change mechanism of Japanese Laid-open Patent Application 2001-349390 in case wherein the range is set from 0.38 to 0.60 without considering the gear spacing ratio. COMPARATIVE EXAMPLES I, II, III and IV are the speed change mechanism shown in FIGS. 9, 10, 13 and 34 of the Laid-open application 2001-349390. 
     FIG. 14  is a comparative graph showing a change of the ratio coverage in case of the speed change mechanism of the invention, the ravigneawx type speed change mechanism and the speed change mechanism of the Laid-open application 2001-349390 with at least one of the planetary gear ratios α 1 , α 2  and α 3  being varied. It is to be noted that in the mechanism of the 2001-349390, the planetary gear ratio α 1  is the ratio possessed by a speed reduction planetary gear unit arranged in corporation with the input shaft, the planetary gear ratio α 3  is the ration possessed by a single pinion type planetary gear unit arranged in corporation with the output shaft, and the planetary gear ratio α 2  is the ratio possessed by a single pinion type planetary gear unit that is arranged between the previously mentioned two planetary gear units. 
   In the following, the ratio coverage of each speed change mechanism will be considered with reference to the contents of the upper row of the table of FIG.  13  and the graph of FIG.  14 . 
   In the ravigneawx type speed change mechanism, the ratio coverage has a narrower range from the minimum value 3.2 to the maximum value 6.7, and thus, widening of the gear ratio coverage is not achieved. Also in the COMPARATIVE EXAMPLES-I, II and III, the ratio coverage has a narrower range, and thus, widening of the gear ratio coverage is not achieved. That is, in such speed change mechanisms, widening of the gear ratio coverage is not expected even if the mechanism has multistage gears. While, in the COMPARATIVE EXAMPLE-IV, the ratio coverage has a wider range from the minimum value 6.1 to the maximum value 15.3, and thus, widening of the gear ratio coverage and increase in freedom in selecting the gear ratio are achieved. 
   In the INVENTION, the ratio coverage has a range from the minimum value 5.1 to the maximum value 9.3. Although the range is somewhat narrower than that of the COMPARATIVE EXAMPLE-IV, satisfied widening of the gear ratio coverage and satisfied increase in freedom in selecting the gear ratio are expected. 
   (4-2) Ratio Coverage with Considering the Gear Spacing Ratio 
   In general, the range of planetary gear ratio “α” is from 0.35 to 0.65. Preferably, the gear spacing ratio is made small as the gear change has a higher speed position. 
   As is seen from the table of  FIG. 12 , in the RAVIGNEAWX-I, the ratio coverage has a narrower range from the minimum value 4.81 to the maximum value 7.20. This is because of a fixed number of teeth of the fifth ring gear R 5  (see FIG.  11 ). 
   While, in the INVENTION-I, the planetary gear ratios “α 1 ” and “α 2 ” of the second and third planetary gear units G 2  and G 3  (see  FIG. 1 ) can be set separately. Accordingly, as is seen from the table of  FIG. 12 , the ratio coverage can have a relatively wider range from the minimum value 4.74 to the maximum value 7.80. Thus, in the invention, increase in freedom in selecting the gear ratio is achieved. 
   The RAVIGNEAWX-I and RAVIGNEAWX-II show their ratio coverage values 6.12 and 6.95 respectively. However, the planetary gear ratios “α 3 ” of the RAVIGNEAWX-I and RAVIGNEAWX-II show values 0.37 and 0.35 respectively. That is, in the ravigneawx type speed change mechanism, when the ratio coverage increases, the planetary gear ratio “α 3 ” becomes small. As has been mentioned hereinabove, reduction of the planetary gear ratio “α 3 ” brings about the radial expansion of the transmission case. 
   In the INVENTION and the RAVIGNEAWX-I, the same values are shown for the gear ratio throughout all of the forward speeds (viz., from 1 st  to 6 th  speeds). However, the planetary gear ratio “α 3 ” (viz., 0.37) of the RAVIGNEAWX-I is quite small as compared with that “α 3 ” (viz., 0.59) of the INVENTION. This means that in the RAVIGNEAWX-I, the transmission case has a radially expanded zone at the portion where the planetary gear unit G 5  is installed (see FIG.  11 ). 
   (5) Comparison in 1-R Ratio 
   In general, the 1-R ratio (viz., gear ratio at the reverse position/gear ratio at 1 st  forward speed) is with a range from 0.8 to 1.2. If the 1-R ratio has a value largely different from such range, output torque produced when an accelerator is pressed becomes different between the reverse position and 1 st  forward speed, which means deterioration in drivability of an associated motor vehicle. 
   The comparative table of  FIG. 13  shows at the lower row respective 1-R ratio values of the speed change mechanism of the invention, the ravigneawx type speed change mechanism of Japanese Laid-open Patent Application 4-219553 and the speed change mechanism of Japanese Laid-open Patent Application 2001-349390. 
     FIG. 15  is a comparative graph showing a change of the 1-R ratio in case of the speed change mechanism of the invention, the ravigneawx type speed change mechanism and the speed change mechanism of the Laid-open application 2001-349390 with at least one of the planetary gear ratios α 1 , α 2  and α 3  being varied. It is to be noted that in the speed change mechanism of the COMPARATIVE EXAMPLE-IV, the range of the 1-R ratio is controlled by varying only the planetary gear ratio α 2 . 
   In the following, the 1-R ratio of each speed change mechanism will be considered with reference to the contents of the lower row of the table of FIG.  13  and the graph of FIG.  15 . 
   In the ravigneawx type speed change mechanism, the two speed change mechanisms of COMPARATIVE EXAMPLE-I and III, the 1-R ratio can have a range from 0.7 to 1.2. That is, the gear ratio for 1 st  forward speed and that for reverse position can have a suitable value. 
   In the speed change mechanism of COMPARATIVE EXAMPLE-II, the 1-R ratio can not have a value lower than 1.22. In this case, the gear ratio at the reverse position become much larger than that at 1 st  forward speed, and thus the drivability of the vehicle becomes deteriorated. 
   In the speed change mechanism of COMPARATIVE EXAMPLE-IV, the 1-R ratio is determined by varying only the planetary gear ratio α 2 . Thus, the 1-R ratio can not have a satisfactorily larger value. That is, as is seen from the graph of  FIG. 15 , in the COMPARATIVE EXAMPLE-IV, a suitable value is not obtained by either of 1 st  forward speed and the reverse position, which means deterioration of drivability of the vehicle. 
   While in the speed change mechanism of the INVENTION, the 1-R ratio can have a range from 0.46 to 0.99, and thus, like the speed change mechanisms of RAVIGNEAWX and COMPARATIVE EXAMPLES-I and III, a suitable gear ratio of the reverse position relative to that for 1 st  forward speed can be set. 
   The entire contents of Japanese Patent Applications 2002-172610 filed Jun. 13, 2002 and 2003-132893 filed May 12, 2003 are incorporated herein by reference. 
   Although the invention has been described above with reference to the embodiment of the invention, the invention is not limited to such embodiment as described above. Various modifications and variations of such embodiment may be carried out by those skilled in the art, in light of the above description.