Planetary gear transmission

A planetary gear transmission comprises a first planetary gear train of double pinion type G1 and second and third planetary gear trains of single pinion type G2 and G3 disposed coaxially and in parallel with one another. In this transmission, a first sun gear S1 is coupled to an input member 11 through a first clutch K1, and a first brake B1 is provided for the purpose of selectively holding the first sun gear S1 against rotation. First and second carriers C1 and C2 and a third ring gear R3 are coupled to one another, and these three elements are coupled to the input member 11 through a second clutch K2, and a second brake B2 is provided for the purpose of selectively holding these three elements against rotation. A first ring gear R1 and a second ring gear R2 are coupled with each other, and a third brake B3 is provided for the purpose of selectively holding these two elements against rotation. Second and third sun gears S2 and S3 are coupled with each other, and these two elements are coupled to the input member 11 through a third clutch K3. A third carrier C3 is directly coupled to an output member 12.

FIELD OF THE INVENTION 
The present invention relates to a planetary gear transmission having three 
planetary gear trains. 
BACKGROUND OF THE INVENTION 
Planetary gear transmissions are widely used as an automatic transmission 
in automobiles. Many conventional planetary gear transmissions comprise an 
integral combination of two planetary gear trains such as the Ravigneaux 
gear train and the Simpson gear train. In addition to such planetary gear 
transmission integrating two gear trains, planetary gear transmissions 
comprising three planetary gear trains are disclosed in Japanese Laid-Open 
Patent Publication No. 1(1989)-320362 and in Japanese Laid-Open (PCT) No. 
4(1992)-501455. In these transmissions, the planetary gear trains are 
combined with a plurality of clutches and brakes for the purpose of 
attaining a plurality of forward gear positions and a single reverse gear 
position. 
When a plurality of forward gear positions, e.g., five or more positions, 
are required; transmissions having three planetary gear trains are popular 
at present because they are effectively constructed with a relatively 
small number of clutches and brakes. For example, the transmission 
disclosed in the above Japanese Laid-Open Patent Publication No. 1-320362 
has three single-planet type planetary gear trains and a total of six 
clutches and brakes for attaining five forward gear positions and a single 
reverse gear position. 
Even though these transmissions comprising three planetary gear trains have 
a small number of clutches and brakes, because of one extra train, they 
tend to have a greater axial length than those comprising only two 
planetary gear trains. However, if the clutches and brakes are arranged 
around the periphery of the planetary gear trains instead of being 
disposed axially next to one another, then the axial length of the 
transmission can be made smaller. 
In the transmission disclosed in the above Japanese Laid-Open Patent 
Publication No. 1-320362, a first ring gear element composing a first 
planetary gear train is coupled with a third ring gear element composing a 
third planetary gear train, and a second ring gear element composing a 
second planetary gear train is coupled to a third carrier element 
composing the third planetary gear train. This construction of the 
transmission requires that a member (first connecting member) connecting 
the first ring gear element with the third ring gear element be disposed 
outside a member (second connecting member) connecting the second ring 
gear element to the third carrier element, thus presenting a problem of 
the first connecting member having a large outside diameter. For the 
purpose of making the axial size of the transmission small, if the 
clutches and brakes were placed around this first connecting member, which 
has a relatively large diameter; then the radial size of the transmission 
would become quite large, which would be another problem. 
In the transmission disclosed in the above Japanese Laid-Open (PCT) No. 
4-501455, the ring gear element of the first planetary gear train is 
coupled to the carrier element of the second planetary gear train, and the 
carrier element of the first planetary gear train is coupled to to the 
ring gear element of the second planetary gear train. This design, which 
disposes connecting members for coupling these elements radially one 
around the other, also presents the above problem. 
SUMMARY OF THE INVENTION 
The present invention is to solve these problems. Thus, it is an object of 
the present invention to provide a small and compact planetary gear 
transmission comprising three planetary gear trains, with at least five 
forward gear positions. 
It is another object of the present invention to provide a planetary gear 
transmission comprising three planetary gear trains, with a design of 
small radial size. 
It is yet another object of the present invention to provide a planetary 
gear transmission comprising three planetary gear trains, with a design of 
small axial size and small radial size. 
In order to achieve these objects, the present invention provides a 
planetary gear transmission having a first planetary gear train of 
double-pinion type, a second planetary gear train of single-pinion type, 
and a third planetary gear train of single-pinion type, each disposed 
coaxially and in parallel with one another. The first planetary gear train 
includes a first sun gear element, a first carrier element, and a first 
ring gear element; the second planetary gear train includes a second sun 
gear element, a second carrier element, and a second ring gear element; 
and the third planetary gear train includes a third sun gear element, a 
third carrier element, and a third ring gear element. The first sun gear 
element is engageably and disengageably coupled to an input member through 
a first clutch, and a first brake is provided capable of holding the first 
sun gear element against rotation. The first and second carrier elements 
and the third ring gear element are coupled to one another, and these 
three elements are engageably and disengageably coupled to the input 
member through a second clutch, and a second brake is provided capable of 
holding these three elements against rotation. Furthermore, the first and 
second ring gear elements are coupled with each other, and a third brake 
is provided capable of holding these two elements against rotation. The 
second and third sun gear elements are coupled with each other, and these 
two elements are engageably and disengageably coupled to the input member 
through a third clutch. In addition, the third carrier element is directly 
coupled to an output member. 
This transmission has the first ring gear element coupled with the second 
ring gear element and the first carrier element with the second carrier 
element such that only the member connecting the first ring gear element 
with the second ring gear element can have a diameter greater than those 
of the first and second ring gear elements in the vicinity of the first 
and second planetary gear trains. Also, the second carrier element is 
coupled to the third ring gear element, and the third carrier element is 
directly connected to the output member. As such, only the member 
connecting the second carrier element to the third ring gear element can 
have a diameter greater than those of the second and third ring gear 
elements in the vicinity of the second and third planetary gear trains. 
Therefore, if the outer diameters of these connecting members are made as 
small as possible and the clutches and brakes are disposed around the 
connecting members (i.e., around the ring gear elements), the transmission 
can be designed with a relatively small length in the axial direction, 
with little compromise of the radial size. 
Furthermore, the first and second ring gear elements can be formed in one 
body by broaching, etc. in one production process, thereby omitting a 
welding process, for producing unified first and second gear elements. 
In the transmission, the first, second, and third planetary gear trains are 
disposed axially adjacent to one another; and the first, second, and third 
clutches and the first brake are disposed on one side of the three 
planetary gear trains in the axial direction, and the second and third 
brakes are disposed on the other side. In addition, the second and third 
clutches are each disposed at a substantially same axial position and are 
overlaid radially one over the other; and the first clutch and the first 
brake are disposed axially adjacent to each other around the second and 
third clutches. Furthermore, the second and third brakes are each disposed 
at a substantially same axial position and are overlaid radially one over 
the other. Preferably, at least the first clutch or the first brake is 
disposed around the first, second, or third ring gear element. In this 
construction, the clutches and brakes are integrally and compactly 
disposed on both sides of the three planetary gear trains. Two clutches 
are disposed radially overlying one over the other, and two brakes are 
also disposed in the same manner. In addition, the first clutch and the 
first brake, or at least one of them, are disposed around the ring gear 
elements. As a result, the transmission has a comparatively small size in 
the axial direction with respect to transmissions having clutches and 
brakes disposed axially adjacent to one another. 
Further scope of applicability of the present invention will become 
apparent from the detailed description given hereinafter. However, it 
should be understood that the detailed description and specific examples, 
while indicating preferred embodiments of the invention, are given by way 
of illustration only, since various changes and modifications within the 
spirit and scope of the invention will become apparent to those skilled in 
the art from this detailed description.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
FIG. 1 shows the whole construction of a planetary gear transmission 
(automatic transmission) according to the present invention, which can be 
incorporated into automobiles. FIG. 2 shows the shift mechanism of the 
transmission in enlargement, and FIG. 3 shows a skeletal diagram of the 
mechanism. 
As shown in FIG. 1, this transmission generally comprises a torque 
converter 5, which is coupled to the output shaft 6 of the engine, and a 
shift mechanism 10 in a three partite housing 1, 2, and 3. The shift 
mechanism 10 transmits power from the input shaft 11 connected to the 
turbines of the torque converter 5 to an output shaft 12 while varying the 
rotational speed ratio. The right-side end of the output shaft 12 
extruding through the right end of the housing 3 is coupled to a propeller 
shaft, which is not shown here. Through this propeller shaft, power is 
transmitted from the out put shaft 12 further to a differential mechanism 
and an axle shaft (i.e., drive wheels), both of which are also not shown 
in the figure. 
As shown in FIG. 2, the input shaft 11 is disposed on the left side of the 
shift mechanism 10; the output shaft 12 is disposed coaxially with the 
input shaft 11 on the right side of the mechanism 10; and a countershaft 
13 is provided coaxially with and between the input and output shafts 11 
and 12. Mounted on this counter shaft 13 are, from the left, first, 
second, and third planetary gear trains G1, G2, and G3, which are disposed 
coaxially, in parallel with and adjacent to one another. The planetary 
gear trains G1, G2, and G3 have respective first, second, and third sun 
gears S1, S2, and S3, which are positioned centrally; respective first, 
second, and third planetary pinions P1, P2, and P3, which mesh with the 
sun gears S1, S2, and S3, respectively, and rotate about their own axes 
while revolving around the sun gears; respective first, second, and third 
carriers C1, C2, and C3, which support the respective planetary pinions 
P1, P2, and P3 rotatably and rotate therewith around the sun gears S1, S2, 
and S3, respectively; and respective first, second, and third ring gears 
R1, R2, and R3, whose internal gear teeth mesh with the planetary pinions 
P1, P2, and P3, respectively. 
Among these planetary gear trains, the second and third planetary gear 
trains G2 and G3 are single-pinion type. However, the first planetary gear 
train G1 is double-pinion type. Therefore, the first pinion P1 has two 
pinion gears P11 and P12. One pinion gear P12 meshes with the first sun 
gear S1, and the other P11 meshes with this pinion gear P12 and the first 
ring gear R1. 
The shift mechanism 10 further comprises first, second, and third clutches 
K1, K2, and K3; first, second, and third brakes B1, B2, and B3; and a 
one-way brake OWB. Each clutch or brake except the one-way brake OWB is 
actuated for engagement when the piston cylinder provided for each is 
supplied with a hydraulic pressure, and it is disengaged by a spring when 
the pressure is released. 
These three clutches K1, K2, and K3 and the first brake B1 are integrally 
disposed on the left side of the three planetary gear trains. On the other 
hand, the second and third brakes B2 and B3 are integrally disposed on the 
right side of the three planetary gear trains. With the clutches and 
brakes placed on both side of the planetary gear trains, the design of the 
transmission is compact. Furthermore, the second and third clutches K2 and 
K3 are each disposed at a substantially same axial position and radially 
overlaid one over the other (the second clutch K2 being outside the third 
clutch K3). Around these clutches K2 and K3, the first clutch K1 and the 
first brake B1 are disposed axially adjacent to each other. In a similar 
manner, the second and third brakes B2 and B3 are each disposed at a 
substantially same axial position and are radially overlaid one over the 
other (the second brake B2 being inside the third brake B3). As a result, 
the shift mechanism 10 (and the whole transmission) has a relatively small 
axial size in comparison with transmissions having clutches and brakes 
disposed axially adjacent to one another. 
The piston mechanism BP1, which is a part of the first brake B1, is 
disposed around the first planetary gear train G1 (i.e., first ring gear 
R1), not axially adjacent to each other. With this positional arrangement, 
the shift mechanism 10 is also axially compacted. In this embodiment of 
the present invention, part of the first brake B1 is disposed around the 
first ring gear R1. However, the whole first brake B1, even together with 
the first clutch K1, can be disposed around the first and second ring 
gears R1 and R2. 
As shown in FIG. 3, the first sun gear S1 is selectively engaged to and 
disengaged from the input shaft 11 by the first clutch K1 or held to the 
housing 2 against rotation by the first brake B1. The first and second 
carriers C1 and C2 and the third ring gear R3 are coupled to one another, 
and they are selectively engaged to and disengaged from the input shaft 11 
by the second clutch K2 or held to housing 3 against rotation by the 
second brake B2. The one-way brake OWB is disposed in parallel with the 
second brake B2, and it is capable of holding the first and second 
carriers C1 and C2 and the third ring gear R3 to the housing 3 against 
rotation in one direction. The first ring gear R1 and the second ring gear 
R2 are coupled with each other, and they can be held to the housing 3 
against rotation by the third brake B3. The second sun gear S2 and the 
third sun gear S3 are coupled with each other, and they are selectively 
engaged to and disengaged from the input shaft 11 by the third clutch K3. 
The third carrier C3 is directly connected to the output shaft 12. 
As shown in FIG. 2, the first and second ring gears R 1 and R2, each having 
a same number of teeth, are formed in one body inside a tubular member 21 
(hereinafter referred to as "first connecting member"). Thus, these ring 
gears R1 and R2 can be formed inside this first connecting member 21 in 
one production process, e.g., by broaching, so it is not necessary that 
these ring gears be separately formed and later be welded together. 
There is nothing between the peripheral surface of the first connecting 
member 21 (i.e., first ring gear R1) and the piston mechanism BP1, which 
is a part of the first brake B1 and disposed around the first connecting 
member 21. Therefore, the first brake B1 is disposed close to the 
peripheral surface of the connecting member 21 so that the transmission 
has a relatively small outer diameter. In this way, with little compromise 
of compactness in the radial direction, the transmission as a whole has a 
relatively small axial size as described previously. 
Furthermore, another tubular member (second connecting member) 22 is 
compactly provided around the third ring gear R3 for the purpose of 
connecting the third ring gear R3 to the second carrier C2. In addition, a 
retaining member 23 is disposed, extending from the first connecting 
member 21 around this second connecting member 22 (i.e., third ring gear 
R3) for the purpose of retaining the third brake B3. With this compact 
arrangement, the radial size of the transmission is not adversely affected 
much. 
As described above, the shift mechanism 10 is composed of the planetary 
gear elements including the first, second, and third sun gears S1, S2, and 
S3, the first, second, and third carriers C1, C2, and C3; the first, 
second, and third ring gears R1, R2, and R3; and the input shaft 11, 
counter shaft 13, and the output shaft 12. In this shift mechanism 10, 
gearshifts or gear positions are controlled and established by selectively 
engaging and disengaging the first, second, and third clutches K1, K2, and 
K3 and the first, second, and third brakes B1, B2, and B3. More 
specifically, five forward gear positions or speed ranges (LOW, 2ND, 3RD, 
4TH, and 5TH) and one reverse gear position or range (REV) are established 
by engaging and disengaging the clutches and brakes in combination as 
shown in FIG. 4. In FIG. 4, the clutches and brakes indicated by a circle 
are engaged. Among these circles, the circle indicating the second brake 
B2 in the LOW range is placed in parentheses. The reason is that power 
will be transmitted through the one-way brake OWB in the LOW range even if 
the second brake B2 is not engaged. In other words, even if the second 
clutch K2 is not engaged, if the third clutch K3 is engaged, then power 
will be transmitted in the LOW range. However, power transmission is not 
possible in the direction opposite to that of the drive side. As such, 
when the second brake B2 is disengaged, engine brake is not available in 
the LOW range. On the other hand, when the second brake B2 is engaged, 
engine brake is operable in the LOW range. 
FIG. 5 is a diagram showing the relation in rotational speed among the 
various elements of the transmission. The speed-reduction ratios of the 
speed ranges will be described below with reference to this diagram. 
In the diagram, the first, second, and third planetary gear trains G1, G2, 
and G3 are separately plotted, and the vertical lines in the respective 
planetary gear trains indicate the elements thereof and have lengths 
representing the rotational speeds of the elements. The distances between 
the vertical lines are proportional to the reciprocals of the numbers of 
teeth of the respective sun gears and ring gears. The clutches K1, K2, and 
K3 and the brakes B1, B2, and B3 are shown adjacent to the respective 
elements, which are selectively engaged thereby. 
For example, the three vertical lines shown in the third planetary gear 
train G3 correspond to the third sun gear S3, the third carrier C3, and 
the third ring gear R3, respectively, from the right. The upward lengths 
of these vertical lines indicate the rotational speeds "n" of these 
elements in the forward direction. The distance "a" between the vertical 
line indicating the third sun gear S3 and the vertical line indicating the 
third carrier C3 corresponds to the reciprocal (=1/Zs) of the number Zs of 
teeth of the third sun gear S3. The distance "b" between the vertical line 
indicating the third carrier C3 and the vertical line indicating the third 
ring gear R3 corresponds to the reciprocal (=1/Zr) of the number Zr of 
teeth of the third ring gear R3. While the third sun gear S3 coupled to 
the input shaft 11 by the engagement of the third clutch K3 rotates at a 
rotational speed "n", if the third ring gear R3 is held against rotation 
by the second brake B2, the third carrier C3 will rotate at a rotational 
speed "nc" which is indicated by the point of intersection between the 
vertical line indicating the third carrier C3 and the line C which 
interconnects the point A indicating the rotating condition of the third 
sun gear S3 and the point B indicating the fixed condition of the third 
ring gear R3. 
The first and second planetary gear trains G1 and G2 are also described in 
the same manner. The three vertical lines in the first planetary gear 
train G1 correspond to the first sun gear S1, the first ring gear R1, and 
the first carrier C1, from the left. 
The first ring gear R1 and the first carrier C1 composing the first 
planetary gear train G1 are coupled with the second ring gear R2 and the 
second carrier C2 composing the second planetary gear train G2, 
respectively. In addition, the second carrier C2 and the second sun gear 
S2 composing the second planetary gear train G2 are coupled to the third 
ring gear R3 and the third sun gear S3 composoing the third planetary gear 
train G3, respectively. Therefore, the diagram in FIG. 5 showing the speed 
relation of the elements in three separate blocks can be redrawn 
integrally as in FIG. 6. With reference to FIG. 6, the rotational speed 
ratio of the output shaft 12 to the input shaft 11, i.e., the 
speed-reduction ratio, will be described for each speed range. 
In the first gear position or LOW range, the third clutch K3 and the second 
brake B2 are engaged (or only the third clutch K3 is engaged as the 
one-way brake OWB is effective). As a result, the second and third sun 
gear S2 and S3 rotate at the same rotational speed n0 as the input shaft 
11, and the first and second carriers C1 and C2 and the third ring gear R3 
are held against rotation. The third carrier C3 rotates at a rotational 
speed n1 which is indicated by the point of intersection between the 
vertical line indicating the third carrier C3 and the dotted line L1 which 
interconnects the point indicating the rotating conditions of the second 
and third sun gears S2 and S3 and the point indicating the fixed 
conditions of the first and second carriers C1 and C2 and the third ring 
gear R3. Thus, the output shaft 12, which is directly connected to the 
third carrier C3, rotates at the rotational speed n1 with a 
speed-reduction ratio decided by the rotational speed ratio of the input 
shaft to the output shaft (=n0/n1). 
In the second gear position, the third clutch K3 and the third brake B3 are 
engaged. The second and third sun gears S2 and S3 rotate at the same 
rotational speed n0 as the input shaft 11, and the first and second ring 
gears R1 and R2 are held against rotation. The output shaft 12 rotates at 
a rotational speed n2 which is indicated by the point of intersection 
between the vertical line indicating the third carrier C3 and the dotted 
line L2 which interconnects the point indicating the rotating conditions 
of the second and third sun gears S2 and S3 and the point indicating the 
fixed conditions of the first and second ring gears R1 and R2. The 
speed-reduction ratio in this gear position is determined by the 
rotational speed ratio of the input shaft to the output shaft (=n0/n2). 
In the third gear position, the third clutch K3 and the first brake B1 are 
engaged. The second and third sun gears S2 and S3 rotate at the same 
rotational speed n0 as the input shaft 11, and the first sun gear S1 is 
held against rotation. The output shaft 12 rotates at a rotational speed 
n3 which is indicated by the point of intersection between the vertical 
line indicating the third carrier C3 and the dotted line L3 which 
interconnects the point indicating the rotating conditions of the second 
and third sun gears S2 and S3 and the point indicating the fixed condition 
of the first sun gear S1. The speed-reduction ratio in this gear position 
is determined by the rotational speed ratio of the input shaft to the 
output shaft (=n0/n3). 
In the fourth gear position, the second and third clutches K2 and K3 are 
engaged. The first, second, and third planetary gear trains G1, G2, and G3 
rotate in unison at the same rotational speed n0 as the input shaft 11, 
and the first sun gears S1 is held against rotation. The output shaft 12 
rotates at a rotational speed n4 which is indicated by the point of 
intersection between the vertical line indicating the third carrier C3 and 
the upper horizontal line L4. The speed-reduction ratio is 1. 
In the fifth gear position, the second clutch K2 and the first brake B1 are 
engaged. The first and second carriers C1 and C2 and the third ring gear 
R3 rotate at the same rotational speed n0 as the input shaft 11, and the 
first sun gear S1 is held against rotation. The output shaft 12 rotates at 
a rotational speed n5 which is indicated by the point of intersection 
between the vertical line indicating the third carrier C3 and the 
extending dotted line L5 which interconnects the point indicating the 
rotating conditions of the first and second carriers C1 and C2 and the 
third ring gear R3 and the point indicating the fixed condition of the 
first sun gear S1. The speed-reduction ratio in this gear position is 
determined by the rotational speed ratio of the input shaft to the output 
shaft (=n0/n5). 
In the reverse gear position or REV range, the first clutch K1 and the 
second brake B2 are engaged. As a result, the first sun gear S1 rotates at 
the same rotational speed n0 as the input shaft 11, and the first and 
second carriers C1 and C2 and the third ring gear R3 are held against 
rotation. The output shaft 12 rotates at a rotational speed nR which is 
indicated by the point of intersection between the vertical line 
indicating the third carrier C3 and the extending dotted line LR which 
interconnects the point indicating the rotating condition of the first sun 
gear S1 and the point indicating the fixed conditions of the first and 
second carriers C1 and C2 and the third ring gear R3. The speed-reduction 
ratio in this gear position is determined by the rotational speed ratio of 
the input shaft to the output shaft (=n0/nR). 
These speed-reduction ratios in the respective speed ranges vary depending 
on the numbers of teeth of the gears, so the values listed as ratios in 
FIG. 4 are given only as a reference. 
The invention being thus described, it will be obvious that the same may be 
varied in many ways. Such variations are not to be regarded as a departure 
from the spirit and scope of the invention, and all such modifications as 
would be obvious to one skilled in the art are intended to be included 
within the scope of the following claims.