Multiple-speed transmission for automotive vehicle

A multiple-speed transmission provides a wide gear ratio coverage for an automotive vehicle by using a basic structure of a six-speed transmission. The multiple-speed transmission is comprised of a combination of two planetary gearsets, three clutches, and three brakes. Each speed is established by applying two of six elements of the three clutches and the three brakes.

BACKGROUND OF THE INVENTION

The present invention relates generally to multiple-speed transmissions for automotive vehicles, and more particularly to a gearbox of a multiple-speed automatic transmission for an automotive vehicle.

Recent years, there have been disclosed various multiple-speed transmissions with more than six forward speeds. One such transmission is disclosed in Japanese Patent Provisional Publication No. 2001-182785 published Jul. 6, 2001. In the publication, a six-speed transmission is comprised of a speed-reducing planetary gearset with one of the rotating members stationary, a Ravigneaux planetary gearset, three clutches, and two brakes. In addition, an eight-speed transmission is comprised by adding a forth clutch to the six-speed transmission.

SUMMARY OF THE INVENTION

In the aforementioned publication, the eight-speed transmission additionally provides a gear ratio between the third speed and the fourth speed, and another gear ratio between the fourth speed and the fifth speed of the six-speed transmission. Accordingly, the gear ratio coverage, that is, the gear ratio of the lowest speed divided by the gear ratio of the highest speed, of the transmission is not varied with the multiplication of gear ratios. However, increasing the gear ratio coverage is desired for improvement of fuel efficiency. With the aforementioned structure, it is difficult to provide wide gear ratio coverage by multiplication of gear ratios with common planetary gearsets.

Accordingly, it is an object of the present invention to provide a multiple-speed transmission for an automotive vehicle that includes a common basic structure and provides a wider gear ratio coverage by multiplication of gear ratios.

In order to accomplish the aforementioned and other objects of the present invention, a multiple-speed transmission for an automotive vehicle comprises an input section, an output section, a speed-reducing planetary gearset including a first rotating element having a first primary speed slower than the input section, a second rotating element drivingly connected to the input section and having a second primary speed faster than the first primary speed in the same direction as the first rotating element, and a third rotating element held against rotation, a speed-shifting planetary gearset including five rotating members including a first rotating member, a second rotating member, a third rotating member, a fourth rotating member, and a fifth rotating member, the second rotating member of the speed-shifting planetary gearset drivingly connected to the output section, a first clutch for selectively drivingly connecting and disconnecting the first rotating element of the speed-reducing planetary gearset and the first rotating member of the speed-shifting planetary gearset, a second clutch for selectively drivingly connecting and disconnecting the first rotating element of the speed-reducing planetary gearset and the fourth rotating member of the speed-shifting planetary gearset, a third clutch for selectively drivingly connecting and disconnecting the second rotating element of the speed-reducing planetary gearset and the third rotating member of the speed-shifting planetary gearset, a first brake operable to hold selectively against rotation the third rotating member of the speed-shifting planetary gearset, a second brake operable to hold selectively against rotation the fourth rotating member of the speed-shifting planetary gearset, and a third brake operable to hold selectively against rotation the fifth rotating member of the speed-shifting planetary gearset.

According to another aspect of the invention, a multiple-speed transmission for an automotive vehicle comprises an input section, an output section, a speed-reducing planetary gearset including a first rotating element having a first primary speed slower than the input section, a second rotating element drivingly connected to the input section and having a second primary speed faster than the first primary speed in the same direction as the first rotating element, and a third rotating element held against rotation, a second planetary gearset including a second sun gear, a third sun gear, a fourth sun gear, a second ring gear drivingly connected to the output section, a second planet-pinion carrier, a second planet pinion rotatably supported on the second planet-pinion carrier and meshed with the third sun gear, a third planet pinion rotatably supported on the second planet-pinion carrier, the third planet pinion comprising a smaller-diameter section meshed with the fourth sun gear, and a larger-diameter section meshed with the second sun gear, the second ring gear, and the second planet pinion, a first clutch for selectively drivingly connecting and disconnecting the first rotating element of the first planetary gearset and the third sun gear, a second clutch for selectively drivingly connecting and disconnecting the first rotating element of the first planetary gearset and the second sun gear, a third clutch for selectively drivingly connecting and disconnecting the second rotating element of the first planetary gearset and the second planet-pinion carrier, a first brake operable to hold selectively against rotation the second planet-pinion carrier, a second brake operable to hold selectively against rotation the second sun gear, and a third brake operable to hold selectively against rotation the fourth sun gear.

According to a further aspect of the invention, a multiple-speed transmission for an automotive vehicle comprises an input section, an output section, a speed-reducing planetary gearset including a first rotating element having a first primary speed slower than the input section, a second rotating element drivingly connected to the input section and having a second primary speed faster than the first primary speed in the same direction as the first rotating element, and a third rotating element held against rotation, a second planetary gearset including a second sun gear, a third sun gear, a second ring gear drivingly connected to the output section, a third ring gear, a second planet-pinion carrier, a second planet pinion rotatably supported on the second planet-pinion carrier and meshed with the third sun gear, a fourth planet pinion rotatably supported on the second planet-pinion carrier and meshed with the third ring gear, a third planet pinion rotatably supported on the second planet-pinion carrier and meshed with the second sun gear, the second ring gear, the second planet pinion, and the fourth planet pinion, a first clutch for selectively drivingly connecting and disconnecting the first rotating element of the first planetary gearset and the third sun gear, a second clutch for selectively drivingly connecting and disconnecting the first rotating element of the first planetary gearset and the second sun gear, a third clutch for selectively drivingly connecting and disconnecting the second rotating element of the first planetary gearset and the second planet-pinion carrier, a first brake operable to hold selectively against rotation the second planet-pinion carrier, a second brake operable to hold selectively against rotation the second sun gear, and a third brake operable to hold selectively against rotation the third ring gear.

According to a still further aspect of the invention, a multiple-speed transmission for an automotive vehicle comprises input means for receiving an input power, output means for outputting a transmitted power, speed-reducing means including a first rotating element having a first primary speed fatie slower than the input means, a second rotating element drivingly connected to the input means and having a second primary speed faster than the first primary speed in the same direction as the first rotating element, and a third rotating element held against rotation, speed-shifting means including five rotating members including a first rotating member, a second rotating member, a third rotating member, a fourth rotating member, and a fifth rotating member, the second rotating member of the speed-shifting planetary gearset drivingly connected to the output means, first torque transmitting means for selectively drivingly connecting and disconnecting the first rotating element of the speed-reducing means and the first rotating member of the speed-shifting means, second torque transmitting means for selectively drivingly connecting and disconnecting the first rotating element of the speed-reducing means and the fourth rotating member of the speed-shifting means, third torque transmitting means for selectively drivingly connecting and disconnecting the second rotating element of the speed-reducing means and the third rotating member of the speed-shifting means, fourth torque transmitting means for holding selectively against rotation the third rotating member of the speed-shifting means, fifth torque transmitting means for holding selectively against rotation the fourth rotating member of the speed-shifting means, and sixth torque transmitting means for holding selectively against rotation the fifth rotating member of the speed-shifting means.

DETAILED DESCRIPTION OF THE INVENTION

Referring now toFIG. 1, there is shown a multiple-speed automatic transmission for an automotive vehicle of a first embodiment of the present invention. The multiple-speed automatic transmission provides eight forward speeds and two reverse speeds. The multiple-speed automatic transmission includes a first planetary gearset G1at the left, and a speed-shifting planetary gearset G2at the right arranged along the axis. First planetary gearset G1is of the single-pinion type, to serve for a speed-reducing planetary gearset as a speed-reducing unit. Speed-shifting planetary gearset G2is a compound planetary gearset.

First planetary gearset G1, which is of the single-pinion type, consists of a first sun gear S1, a first ring gear R1, and a first planet-pinion carrier PC1that carries or rotatably supports a first planet pinion P1meshed with both first sun gear S1and first ring gear R1. First sun gear S1as a third rotating element e3is continuously held against rotation to a transmission housing. First planet-pinion carrier PC1serves for a first rotating element e1of first planetary gearset G1with a first primary speed slower than that of first ring gear R1that serves for a second rotating element e2of first planetary gearset G1.

Speed-shifting planetary gearset G2consists of three sun gears of a second sun gear S2, a third sun gear S3and a fourth sun gear S4, a second ring gear R2, and a second planet-pinion carrier PC2that carries or rotatably supports two planet pinions of a second planet pinion Pb1and a third planet pinion Pb2. Third planet pinion Pb2is formed of a two-stepped cylindrical shape, that is, includes integrally two sections of a smaller-diameter section Pb2aand a larger-diameter section Pb2bthat have two different numbers of teeth. Smaller-diameter section Pb2aof third planet pinion Pb2is meshed with fourth sun gear S4. Larger-diameter section Pb2bof third planet pinion Pb2is meshed with second ring gear R2, second sun gear S2and second planet pinion Pb1. Second planet pinion Pb1is meshed with third sun gear S3.

Speed-shifting planetary gearset G2includes five rotating members. A first rotating member m1consists of elements that rotate solidly with third sun gear S3. A second rotating member m2consists of elements that rotate solidly with second ring gear R2. A third rotating member m3consists of elements that rotate solidly with second planet-pinion carrier PC2. A fourth rotating member m4consists of elements that rotate solidly with second sun gear S2. A fifth rotating member m5consists of elements that rotate solidly with fourth sun gear S4.

The aforementioned structure is connected to an input section, such as an input shaft IP1and an output section, such as an output shaft OP1. Input shaft IP1is drivingly connected to first ring gear R1, to input driving torque transmitted via a torque converter (not shown) and others from an engine (not shown) as a drive source. Output shaft OP1is drivingly connected to second ring gear R2, to output driving torque via a final gear (not shown) and others to a driving wheel (not shown).

Additionally, the multiple-speed automatic transmission includes three clutches and three brakes. A first clutch C1selectively connects or disconnects first planet-pinion carrier PC1and third sun gear S3(first rotating member m1). A second clutch C2selectively connects or disconnects first planet-pinion carrier PC1and second sun gear S2(fourth rotating member m4). A third clutch C3selectively connects or disconnects input shaft IP1and second planet-pinion carrier PC2(third rotating member m3). A first brake B1is operable to selectively hold against rotation to the transmission housing or release second planet-pinion carrier PC2(third rotating member m3). A second brake B2is operable to selectively hold against rotation to the transmission housing or release second sun gear S2(fourth rotating member m4). A third brake B3is operable to selectively hold against rotation to the transmission housing or release fourth sun gear S4(fifth rotating member m5).

Clutches C1, C2, and C3and brakes B1, B2, and B3are connected to a transmission control unit (not shown) as a transmission controlling means for supplying engaging pressure and releasing pressure according to clutch engagements and brake applications required to establish various gear speeds. The transmission control unit may be of the hydraulic control type, the electronic control type, or the electrohydraulic control type.

Referring now toFIGS. 2 through 7B, the following describes the operation of the multiple-speed automatic transmission of the first embodiment.FIG. 2shows clutch engagements and brake applications required to establish various gear speeds. InFIG. 2, a solid circle in a cell indicates that the corresponding clutch or brake is applied in the corresponding speed, and a blank indicates that the corresponding clutch or brake is released in the corresponding speed.FIG. 3shows the collinear diagram of the multiple-speed automatic transmission. The collinear diagram shows the rotation states of the rotating members in each speed. InFIG. 3, a bold line indicates the collinear diagram of first planetary gearset G1, and medium bold lines indicate the collinear diagram of speed-shifting planetary gearset G2. Speed-shifting planetary gearset G2takes a rotation state determined by a combination of rotations of two of the five rotating members, where each of the five rotating members of speed-shifting planetary gearset G2has a rotation speed that monotonously varies in order of first rotating member m1, second rotating member m2, third rotating member m3, fifth rotating member m5, and fourth rotating member m4.FIGS. 4A to 7Bshow the power flow or the torque flow in each speed. InFIGS. 4A to 7B, the power flow through the clutches, the brakes, and the rotating members is indicated by bold lines, the power flow through the gears is indicated by a hatch pattern.

The first speed is established by engaging first clutch C1and applying first brake B1, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. At speed-shifting planetary gearset G2, with first clutch C1engaged, the reduced speed is input to third sun gear S3from first planet-pinion carrier PC1of first planetary gearset G1. On the other hand, with first brake B1applied, second planet-pinion carrier PC2is fixed to the transmission housing. In this state, third sun gear S3, second planet-pinion carrier PC2that carries second planet pinion Pb1and larger-diameter section Pb2bof third planet pinion Pb2, and second ring gear R2are assumed to form a planetary gearset of the double-pinion type with a planet-pinion carrier fixed. Therefore, with third sun gear S3rotating at a speed reduced at first planetary gearset G1, second ring gear R2is forced to rotate at a further reduced speed, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 3provides another analysis of the first speed. The point identified by C1in the diagram, or the engagement point of first clutch C1indicates the engagement of first clutch C1with which the reduced speed is input to third sun gear S3from first planetary gearset G1. The point identified by B1in the diagram, or the application point of first brake B1indicates the application of first brake B1with which second planet-pinion carrier PC2is held stationary. The operation of speed-shifting planetary gearset G2is defined by the lever or straight line connecting the engagement point of first clutch C1and the application point of first brake B1. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the first speed, the rotation speed of input shaft IP1is reduced to a point identified by 1ST in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the first speed is shown inFIG. 4A. The power flows through first clutch C1, first brake B1, and the rotating members, as shown by bold lines, and first planetary gearset G1, and speed-shifting planetary gearset G2except second sun gear S2and fourth sun gear S4, as shown by a hatch pattern. In this speed, first planetary gearset G1and speed-shifting planetary gearset G2serve for the transmission of power and torque.

The second speed is established by releasing first brake B1and applying third brake B3to the operational state of the first speed, that is, by engaging first clutch C1and applying third brake B3, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. At speed-shifting planetary gearset G2, with first clutch C1engaged, the reduced speed is input to third sun gear S3from first planet-pinion carrier PC1of first planetary gearset G1. On the other hand, with third brake B3applied, fourth sun gear S4is fixed to the transmission housing. Fourth sun gear S4, second planet-pinion carrier PC2that carries third planet pinion Pb2, and second ring gear R2are assumed to form a planetary gearset. With fourth sun gear S4fixed and second ring gear R2rotating, second planet-pinion carrier PC2is forced to rotate at a speed reduced from the speed of second ring gear R2. At the same time, third sun gear S3, second planet-pinion carrier PC2that carries second planet pinion Pb1and third planet pinion Pb2, and second ring gear R2are assumed to form a planetary gearset of the double-pinion type with a planet-pinion carrier rotating at a low speed. Therefore, with third sun gear S3rotating at a speed reduced at first planetary gearset G1, second ring gear R2is forced to rotate at a further reduced speed (faster than the first speed), thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 3provides another analysis of the second speed. The point identified by C1in the diagram, or the engagement point of first clutch C1indicates the engagement of first clutch C1with which the reduced speed is input to third sun gear S3from first planetary gearset G1. The point identified by B3in the diagram, or the application point of third brake B3indicates the application of third brake B3with which fourth sun gear S4is held stationary. The operation of speed-shifting planetary gearset G2is defined by the lever or straight line connecting the engagement point of first clutch C1and the application point of third brake B3. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the second speed, the rotation speed of input shaft IP1is reduced to a point identified by 2ND in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the second speed is shown inFIG. 4B. The power flows through first clutch C1, third brake B3, and the rotating members, as shown by bold lines, first planetary gearset G1and speed-shifting planetary gearset G2except second sun gear S2, as shown by a hatch pattern. In this speed, first planetary gearset G1and speed-shifting planetary gearset G2serve for the transmission of power and torque.

The third speed is established by releasing third brake B3and applying second brake B2to the operational state of the second speed, that is, by engaging first clutch C1and applying second brake B2, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. At speed-shifting planetary gearset G2, with first clutch C1engaged, the reduced speed is input to third sun gear S3from first planet-pinion carrier PC1of first planetary gearset G1. On the other hand, with second brake B2applied, second sun gear S2is fixed to the transmission housing. Second sun gear S2, second planet-pinion carrier PC2that carries larger-diameter section Pb2bof third planet pinion Pb2, and second ring gear R2are assumed to form a planetary gearset. With second sun gear S2fixed and second ring gear R2rotating, second planet-pinion carrier PC2is forced to rotate at a speed reduced from the speed of second ring gear R2. At the same time, third sun gear S3, second planet-pinion carrier PC2that carries second planet pinion Pb1and third planet pinion Pb2, and second ring gear R2are assumed to form a planetary gearset of the double-pinion type with a planet-pinion carrier rotating at a low speed. Therefore, with third sun gear S3rotating at a speed reduced at first planetary gearset G1, second ring gear R2is forced to rotate at a further reduced speed that is faster than the second speed according to the difference between gear ratios of second sun gear S2to larger-diameter section Pb2bof third planet pinion Pb2and fourth sun gear S4to smaller-diameter section Pb2aof third planet pinion Pb2, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 3provides another analysis of the third speed. The point identified by C1in the diagram, or the engagement point of first clutch C1indicates the engagement of first clutch C1with which the reduced speed is input to third sun gear S3from first planetary gearset G1. The point identified by B2in the diagram, or the application point of second brake B2indicates the application of second brake B2with which second sun gear S2is held stationary. The operation of speed-shifting planetary gearset G2is defined by the lever or straight line connecting the engagement point of first clutch C1and the application point of second brake B2. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the third speed, the rotation speed of input shaft IP1is reduced to a point identified by 3RD in the diagram (faster than the second speed) through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the second speed is shown inFIG. 4C. The power flows through first clutch C1, second brake B2, and the rotating members, as shown by bold lines, and first planetary gearset G1and speed-shifting planetary gearset G2except fourth sun gear S4, as shown by a hatch pattern.

The fourth speed is established by releasing second brake B2and engaging second clutch C2to the operational state of the third speed, that is, by engaging first clutch C1and second clutch C2, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. At speed-shifting planetary gearset G2, with first clutch C1engaged, the reduced speed is input to third sun gear S3from first planet-pinion carrier PC1of first planetary gearset G1. At the same time, with second clutch C2engaged, the reduced speed is input to second sun gear S2from first planet-pinion carrier PC1of first planetary gearset G1. Accordingly, at speed-shifting planetary gearset G2, third sun gear S3and second sun gear S2rotate at the same reduced speed, so that second planet-pinion carrier PC2and second ring gear R2also rotate solidly with third sun gear S3and second sun gear S2. Therefore second ring gear R2is forced to rotate at the reduced speed that is reduced at first planetary gearset G1, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 3provides another analysis of the fourth speed. The point identified by C1in the diagram, or the engagement point of first clutch C1indicates the engagement of first clutch C1with which the reduced speed is input to third sun gear S3from first planetary gearset G1. Similarly, the point identified by C2in the diagram, or the engagement point of second clutch C2indicates the engagement of second clutch C2with which the reduced speed is input to second sun gear S2from first planetary gearset G1. The operation of speed-shifting planetary gearset G2is defined by the lever or straight line connecting the engagement point of first clutch C1and the engagement point of second clutch C2. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the fourth speed, the rotation speed of input shaft IP1is reduced to a point identified by 4TH in the diagram (to the gear ratio of first planetary gearset G1) through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the second speed is shown inFIG. 5A. The power flows through first clutch C1, second clutch C2, and the rotating members, as shown by bold lines, and first planetary gearset G1and speed-shifting planetary gearset G2except fourth sun gear S4, as shown by a hatch pattern.

The fifth speed is established by disengaging second clutch C2and engaging third clutch C3to the operational state of the fourth speed, that is, by engaging first clutch C1and third clutch C3, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. At speed-shifting planetary gearset G2, with first clutch C1engaged, the reduced speed is input to third sun gear S3from first planet-pinion carrier PC1of first planetary gearset G1. On the other hand, with third clutch C3engaged, the input rotation of input shaft IP1is input to second planet-pinion carrier PC2. In this state, third sun gear S3, second planet-pinion carrier PC2that carries second planet pinion Pb1and larger-diameter section Pb2bof third planet pinion Pb2, and second ring gear R2are assumed to form a planetary gearset of the double-pinion type. Therefore, with third sun gear S3rotating at a speed reduced at first planetary gearset G1and second planet-pinion carrier PC2rotating at the input speed, second ring gear R2is forced to rotate at a speed intermediate between that of third sun gear S3and that of second planet-pinion carrier PC2, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 3provides another analysis of the fifth speed. The point identified by C1in the diagram, or the engagement point of first clutch C1indicates the engagement of first clutch C1with which the reduced speed is input to third sun gear S3from first planetary gearset G1. Similarly, the point identified by C3in the diagram, or the engagement point of third clutch C3indicates the engagement of third clutch C3with which the input speed is input to second planet-pinion carrier PC2from input shaft IP1. The operation of speed-shifting planetary gearset G2is defined by the lever or straight line connecting the engagement point of first clutch C1and the engagement point of third clutch C3. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the fifth speed, the rotation speed of input shaft IP1is reduced slightly to a point identified by 5TH in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the fifth speed is shown inFIG. 5B. The power flows through first clutch C1, third clutch C3, and the rotating members, as shown by bold lines, and first planetary gearset G1and speed-shifting planetary gearset G2except second sun gear S2and fourth sun gear S4, as shown by a hatch pattern.

The sixth speed is established by disengaging first clutch C1and engaging second clutch C2to the operational state of the fifth speed, that is, by engaging second clutch C2and third clutch C3, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. With second clutch C2engaged, the reduced speed is input to second sun gear S2from first planet-pinion carrier PC1of first planetary gearset G1. At the same time, with third clutch C3engaged, the input rotation of input shaft IP1is input to second planet-pinion carrier PC2. Accordingly, in this state, second sun gear S2, second planet-pinion carrier PC2that carries larger-diameter section Pb2bof third planet pinion Pb2, and second ring gear R2are assumed to form a planetary gearset of the single-pinion type. Therefore, with second sun gear S2rotating at a speed reduced at first planetary gearset G1and second planet-pinion carrier PC2rotating at the input speed, second ring gear R2is forced to rotate at a speed increased from the input speed, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 3provides another analysis of the sixth speed. The point identified by C2in the diagram, or the engagement point of second clutch C2indicates the engagement of second clutch C2with which the reduced speed is input to second sun gear S2from first planetary gearset G1. Similarly, the point identified by C3in the diagram, or the engagement point of third clutch C3indicates the engagement of third clutch C3with which the input speed is input to second planet-pinion carrier PC2from input shaft IP1. The operation of speed-shifting planetary gearset G2is defined by the lever or straight line connecting the engagement point of second clutch C2and the engagement point of third clutch C3. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the sixth speed, the rotation speed of input shaft IP1is increased slightly to a point identified by 6TH in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the sixth speed is shown inFIG. 5C. The power flows through second clutch C2, third clutch C3, and the rotating members, as shown by bold lines, and first planetary gearset G1and speed-shifting planetary gearset G2except third sun gear S3, fourth sun gear S4, and second planet pinion Pb1, as shown by a hatch pattern.

The seventh speed is established by disengaging second clutch C2and applying second brake B2to the operational state of the sixth speed, that is, by engaging third clutch C3and applying second brake B2, as shown inFIG. 2. With third clutch C3engaged, the input rotation speed of input shaft IP1is input to second planet-pinion carrier PC2of speed-shifting planetary gearset G2. On the other hand, with second brake B2applied, second sun gear S2of speed-shifting planetary gearset G2is held stationary to the transmission housing. In this state, second sun gear S2, second planet-pinion carrier PC2that carries larger-diameter section Pb2bof third planet pinion Pb2, and second ring gear R2are assumed to form a planetary gearset of the single-pinion type with a sun gear fixed. Therefore, with second planet-pinion carrier PC2rotating at the input speed, second ring gear R2is forced to rotate at a speed increased faster than the input speed, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 3provides another analysis of the seventh speed. The point identified by C3in the diagram, or the engagement point of third clutch C3indicates the engagement of third clutch C3with which the input speed is input to second planet-pinion carrier PC2from input shaft IP1. The point identified by B2in the diagram, or the application point of second brake B2indicates the application of second brake B2with which second sun gear S2is held stationary. The operation of speed-shifting planetary gearset G2is defined by the lever or straight line connecting the engagement point of third clutch C3and the application point of second brake B2. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the seventh speed, the rotation speed of input shaft IP1is increased to a point identified by 7TH in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the seventh speed is shown inFIG. 6A. The power flows through third clutch C3, second brake B2, and the rotating members, as shown by bold lines, and speed-shifting planetary gearset G2except third sun gear S3, fourth sun gear S4, and second planet pinion Pb1, as shown by a hatch pattern.

The eighth speed is established by releasing second brake B2and applying third brake B3to the operational state of the seventh speed, that is, by engaging third clutch C3and applying third brake B3, as shown inFIG. 2. With third clutch C3engaged, the input rotation speed of input shaft IP1is input to second planet-pinion carrier PC2of speed-shifting planetary gearset G2. On the other hand, with third brake B3applied, fourth sun gear S4of speed-shifting planetary gearset G2is held stationary to the transmission housing. Fourth sun gear S4, second planet-pinion carrier PC2that carries third planet pinion Pb2, and second ring gear R2are assumed to form a planetary gearset of the single-pinion type with a sun gear fixed. Therefore, with second planet-pinion carrier PC2rotating at the input speed, second ring gear R2is forced to rotate at a speed increased from the input speed that is faster than the seventh speed according to the difference between gear ratios of second sun gear S2to larger-diameter section Pb2bof third planet pinion Pb2and fourth sun gear S4to smaller-diameter section Pb2aof third planet pinion Pb2since fourth sun gear S4is meshed with smaller-diameter section Pb2aof third planet pinion Pb2.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 3provides another analysis of the eighth speed. The point identified by C3in the diagram, or the engagement point of third clutch C3indicates the engagement of third clutch C3with which the input speed is input to second planet-pinion carrier PC2from input shaft IP1. The point identified by B3in the diagram, or the application point of third brake B3indicates the application of third brake B3with which fourth sun gear S4is held stationary to the transmission housing. The operation of speed-shifting planetary gearset G2is defined by the lever or straight line connecting the engagement point of third clutch C3and the application point of third brake B3. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the eighth speed, the rotation speed of input shaft IP1is increased to a point identified by 8TH in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the eighth speed is shown inFIG. 6B. The power flows through third clutch C3, third brake B3, and the rotating members, as shown by bold lines, and speed-shifting planetary gearset G2except second sun gear S2, third sun gear S3, and second planet pinion Pb1, as shown by a hatch pattern.

The first reverse speed is established by engaging second clutch C2and applying first brake B1, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. With second clutch C2engaged, the reduced speed is input to second sun gear S2from first planet-pinion carrier PC1of first planetary gearset G1. On the other hand, with first brake B1applied, second planet-pinion carrier PC2is fixed to the transmission housing. In this state, second sun gear S2, second planet-pinion carrier PC2that carries larger-diameter section Pb2bof third planet pinion Pb2, and second ring gear R2are assumed to form a planetary gearset of the single-pinion type with a planet-pinion carrier fixed. Therefore, with second sun gear S2rotating at the reduced speed, second ring gear R2is forced to rotate in the reverse direction at a reduced speed, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 3provides another analysis of the first reverse speed. The point identified by C2in the diagram, or the engagement point of second clutch C2indicates the engagement of second clutch C2with which the reduced speed is input to second sun gear S2from first planetary gearset G1. The point identified by B1in the diagram, or the application point of first brake B1indicates the application of first brake B1with which second planet-pinion carrier PC2is held stationary. The operation of speed-shifting planetary gearset G2is defined by the lever or straight line connecting the engagement point of second clutch C2and the application point of first brake B1. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the first reverse speed, the rotation speed of input shaft IP1is reduced to a point identified by REV1 in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the first reverse speed is shown inFIG. 7A. The power flows through second clutch C2, first brake B1; and the rotating members, as shown by bold lines, and first planetary gearset G1and speed-shifting planetary gearset G2except third sun gear S3, fourth sun gear S4, and second planet pinion Pb1, as shown by a hatch pattern.

The second reverse speed is established by releasing first brake B1and applying third brake B3to the operational state of the first reverse speed, that is, by engaging second clutch C2and applying third brake B3, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. With second clutch C2engaged, the reduced speed is input to second sun gear S2from first planet-pinion carrier PC1of first planetary gearset GI. On the other hand, with third brake B3applied, fourth sun gear S4is fixed to the transmission housing. Fourth sun gear S4, second planet-pinion carrier PC2that carries third planet pinion Pb2, and second ring gear R2are assumed to form a planetary gearset. With fourth sun gear S4fixed and second ring gear R2rotating at the reduced speed, second planet-pinion carrier PC2is forced to rotate at a speed reduced from the speed of second ring gear R2. At the same time, second sun gear S2, second planet-pinion carrier PC2that carries larger-diameter section Pb2bof third planet pinion Pb2, and second ring gear R2are assumed to form a planetary gearset of the single-pinion type with a planet-pinion carrier rotating at a low speed. Therefore, with second sun gear S2rotating at a speed reduced at first planetary gearset G1, second ring gear R2is forced to rotate in the reverse direction at a reduced speed (faster than the first reverse speed), thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 3provides another analysis of the second reverse speed. The point identified by C2in the diagram, or the engagement point of second clutch C2indicates the engagement of second clutch C2with which the reduced speed is input to second sun gear S2from first planetary gearset G1. The point identified by B3in the diagram, or the application point of third brake B3indicates the application of third brake B3with which fourth sun gear S4is held stationary. The operation of speed-shifting planetary gearset G2is defined by the lever or straight line connecting the engagement point of second clutch C2and the application point of third brake B3. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the second reverse speed, the rotation speed of input shaft IP1is reduced to a point identified by REV2 in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the second reverse speed is shown inFIG. 7B. The power flows through second clutch C2, third brake B3, and the rotating members, as shown by bold lines, and first planetary gearset G1and speed-shifting planetary gearset G2except third sun gear S3and second planet pinion Pb1, as shown by a hatch pattern.

The multiple-speed automatic transmission of the first embodiment is constructed and operated as discussed above. The following describes the comparison between the multiple-speed automatic transmission of the first embodiment and the corresponding six-speed automatic transmission that serves for the basic structure of the multiple-speed automatic transmission of the first embodiment. Referring now toFIG. 14, there is shown the six-speed automatic transmission. Elements in common are designated by the same reference signs. The six-speed automatic transmission includes two planetary gearsets of first planetary gearset G1and speed-shifting planetary gearset G22, three clutches C1through C3, and two brakes B1and B2. The difference is that the multiple-speed automatic transmission includes additional elements of fourth sun gear S4and smaller-diameter section Pb2aof third planet pinion Pb2meshed with fourth sun gear S4, and third brake B3. In other words, simply adding fourth sun gear S4, smaller-diameter section Pb2ain third planet pinion Pb22, and third brake B3serves for multiplication of gear speeds from the six-speed automatic transmission to the eight-speed automatic transmission of the first embodiment.

The following describes effects of the first embodiment. (E1) The eight-speed automatic transmission of the first embodiment is comprised by fourth sun gear S4, smaller-diameter section Pb2ain third planet pinion Pb22, and third brake B3to the corresponding six-speed automatic transmission. As shown inFIG. 2, a gear speed between the first speed and the second speed of the corresponding six-speed automatic transmission is provided by engaging first clutch C1and applying third brake B3. In addition, a gear speed faster than the sixth speed of the corresponding six-speed automatic transmission is provided by engaging third clutch C3and applying third brake B3. In summary, an eight-speed automatic transmission can be provided by applying planetary gearsets with the same gear ratio as the corresponding six-speed automatic transmission, while the gear ratios can be properly set. Addition of the top speed or the eighth speed provides a wider ratio coverage, to improve fuel efficiency, which is an object of multiplication of gear speeds. As shown inFIG. 2, for example, the ratio coverage of the automatic transmission of the first embodiment is increased to 8.13, while the ratio coverage of the six-speed automatic transmission is equal to 6.35. This is derived from the gear ratios of the planetary gearsets. In this example, the gear ratios of α1or first sun gear S1/first ring gear R1gear ratio, α2or second sun gear S2/second ring gear R2gear ratio, α3or third sun gear S3/second ring gear R2, and α4or fourth sun gear S4/second ring gear R2gear ratio are equal to 0.58, 0.4, 0.35, and 0.60, respectively. In addition, the numbers of teeth of second sun gear S2, larger-diameter section Pb2b, second ring gear R2, fourth sun gear S4, and smaller-diameter section Pb2aare set to 40, 40, 100, 60, and 30, respectively.

(E2) The new second speed is added between the first speed and the second speed of the corresponding six-speed automatic transmission, where torque difference is large, to reduce torque difference and shift shock. As shown inFIG. 2, the difference of gear ratio between the first speed and the second speed of the eight-speed automatic transmission of the first embodiment is equal to 1.63, and the difference between the second and the third is equal to 0.46, while the difference of gear ratio between the first speed and the second speed of the corresponding six-speed automatic transmission is equal to 2.09. Accordingly, this provides smoother shifting from the first speed to the second speed. On the other hand, addition of the new second speed leads to lowering the minimum speed for lock-up, to improve fuel efficiency. In general, the first speed is used during starting from a stationary state, without lock-up, from the viewpoint of fail-safe operation. In other words, decreasing the second speed that usually determines the minimum speed for lock-up, or increasing the gear ratio of the second leads to a low minimum speed for lock-up and improvement of fuel efficiency.

(E3) An additional reverse speed faster than the reverse speed of the corresponding six-speed automatic transmission is provided. The new reverse speed is established by engaging second clutch C2and applying third brake B3, as discussed above.

Referring now toFIG. 8, there is shown a multiple-speed automatic transmission for an automotive vehicle of a second embodiment of the present invention. The multiple-speed automatic transmission provides eight forward speeds and two reverse speeds. The multiple-speed automatic transmission includes a first planetary gearset G1at the left, and a speed-shifting planetary gearset G21at the right arranged along the axis. First planetary gearset G1is of the single-pinion type, to serve for a speed-reducing planetary gearset as a speed-reducing unit. Speed-shifting planetary gearset G21is a compound planetary gearset.

First planetary gearset G1, which is of the single-pinion type, consists of a first sun gear S1, a first ring gear R1, and a first planet-pinion carrier PC1that carries or rotatably supports a first planet pinion P1meshed with both first sun gear S1and first ring gear R1. First sun gear S1as a third rotating element e3is continuously held against rotation to a transmission housing. First planet-pinion carrier PCI serves for a first rotating element e1of first planetary gearset G1with a first primary speed slower than that of first ring gear R1that serves for a second rotating element e2of first planetary gearset G1.

Speed-shifting planetary gearset G21consists of two sun gears of a second sun gear S2and a third sun gear S3, two ring gears of a second ring gear R2and a third ring gear R3, and a second planet-pinion carrier PC2that carries or rotatably supports three planet pinions of a second planet pinion Pb1, a third planet pinion Pb21, and a fourth planet pinion Pb3. Second planet pinion Pb1is meshed with third sun gear S3and third planet pinion Pb21. Third planet pinion Pb21is meshed with second sun gear S2, second ring gear R2, and fourth planet pinion Pb3. Fourth planet pinion Pb3is meshed with third ring gear R3. Third planet pinion Pb21in the second embodiment has a single outside diameter and a single number of teeth, while third planet pinion Pb2in the first embodiment is formed of a two-stepped cylindrical shape.

Speed-shifting planetary gearset G21includes five rotating members. A first rotating member m1consists of elements that rotate solidly with third sun gear S3. A second rotating member m2consists of elements that rotate solidly with second ring gear R2. A third rotating member m3consists of elements that rotate solidly with second planet-pinion carrier PC2. A fourth rotating member m4consists of elements that rotate solidly with second sun gear S2. A fifth rotating member m5consists of elements that rotate solidly with third ring gear R3.

The aforementioned structure is connected to an input section, such as an input shaft IP1and an output section, such as an output shaft OP1. Input shaft IP1is drivingly connected to first ring gear R1, to input driving torque transmitted via a torque converter (not shown) and others from an engine (not shown) as a drive source. Output shaft OP1is drivingly connected to second ring gear R2, to output driving torque via a final gear (not shown) and others to a driving wheel (not shown).

Additionally, the multiple-speed automatic transmission includes three clutches and three brakes. A first clutch C1selectively connects or disconnects first planet-pinion carrier PC1and third sun gear S3(first rotating member m1). A second clutch C2selectively connects or disconnects first planet-pinion carrier PC1and second sun gear S2(fourth rotating member m4). A third clutch C3selectively connects or disconnects input shaft IP1and second planet-pinion carrier PC2(third rotating member m3). A first brake B1is operable to selectively hold against rotation to the transmission housing or release second planet-pinion carrier PC2(third rotating member m3). A second brake B2is operable to selectively hold against rotation to the transmission housing or release second sun gear S2(fourth rotating member m4). A third brake B3is operable to selectively hold against rotation to the transmission housing or release third ring gear R3(fifth rotating member m5).

Referring now toFIGS. 2 and 9through13B, the following describes the operation of the multiple-speed automatic transmission of the first embodiment.FIG. 2, partly shared between the first and second embodiments, shows clutch engagements and brake applications required to establish various gear speeds. InFIG. 2, a solid circle in a cell indicates that the corresponding clutch or brake is applied in the corresponding speed, and a blank indicates that the corresponding clutch or brake is released in the corresponding speed.FIG. 9shows the collinear diagram of the multiple-speed automatic transmission. The collinear diagram shows the rotation states of the rotating members in each speed. InFIG. 9, a bold line indicates the collinear diagram of first planetary gearset G1, and medium bold lines indicate the collinear diagram of speed-shifting planetary gearset G21. Speed-shifting planetary gearset G21gearset takes a rotation state determined by a combination of rotations of two of the five rotating members, where each of the five rotating members of speed-shifting planetary gearset G21has a rotation speed that monotonously varies in order of first rotating member m1, second rotating member m2, third rotating member m3, fifth rotating member m5, and fourth rotating member m4.FIGS. 10A to 13Bshow the power flow or the torque flow in each speed. InFIGS. 10A to 13B, the power flow through the clutches, the brakes, and the rotating members is indicated by bold lines, the power flow through the gears is indicated by a hatch pattern.

The first speed is established by engaging first clutch C1and applying first brake B1, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. At speed-shifting planetary gearset G21, with first clutch C1engaged, the reduced speed is input to third sun gear S3from first planet-pinion carrier PC1of first planetary gearset G1. On the other hand, with first brake B1applied, second planet-pinion carrier PC2is fixed to the transmission housing. In this state, third sun gear S3, second planet-pinion carrier PC2that carries second planet pinion Pb1and third planet pinion Pb21, and second ring gear R2are assumed to form a planetary gearset of the double-pinion type with a planet-pinion carrier fixed. Therefore, with third sun gear S3rotating at a speed reduced at first planetary gearset G1, second ring gear R2is forced to rotate at a further reduced speed, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 9provides another analysis of the first speed. The point identified by C1in the diagram, or the engagement point of first clutch C1indicates the engagement of first clutch C1with which the reduced speed is input to third sun gear S3from first planetary gearset G1. The point identified by B1in the diagram, or the application point of first brake B1indicates the application of first brake B1with which second planet-pinion carrier PC2is held stationary. The operation of speed-shifting planetary gearset G21is defined by the lever or straight line connecting the engagement point of first clutch C1and the application point of first brake B1. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the first speed, the rotation speed of input shaft IP1is reduced to a point identified by 1ST in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the first speed is shown inFIG. 10A. The power flows through first clutch C1, first brake B1, and the rotating members, as shown by bold lines, and first planetary gearset G1, and speed-shifting planetary gearset G21except second sun gear S2, fourth planet pinion Pb3, and third ring gear R3, as shown by a hatch pattern.

The second speed is established by releasing first brake B1and applying third brake B3to the operational state of the first speed, that is, by engaging first clutch C1and applying third brake B3, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. At speed-shifting planetary gearset G21, with first clutch C1engaged, the reduced speed is input to third sun gear S3from first planet-pinion carrier PC1of first planetary gearset G1. On the other hand, with third brake B3applied, third ring gear R3is fixed to the transmission housing. With third ring gear R3fixed and third sun gear S3rotating at the reduced speed, second planet-pinion carrier PC2is forced to rotate at a speed further decreased from that of third sun gear S3. At the same time, third sun gear S3, second planet-pinion carrier PC2that carries second planet pinion Pb1and third planet pinion Pb21, and second ring gear R2are assumed to form a planetary gearset of the double-pinion type with a planet-pinion carrier rotating at a low speed. Therefore, with third sun gear S3rotating at a speed reduced at first planetary gearset G1, second ring gear R2is forced to rotate at a further reduced speed (faster than the first speed), thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 9provides another analysis of the second speed. The point identified by C1in the diagram, or the engagement point of first clutch C1indicates the engagement of first clutch C1with which the reduced speed is input to third sun gear S3from first planetary gearset G1. The point identified by B3in the diagram, or the application point of third brake B3indicates the application of third brake B3with which third ring gear R3is held stationary. The operation of speed-shifting planetary gearset G21is defined by the lever or straight line connecting the engagement point of first clutch C1and the application point of third brake B3. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the second speed, the rotation speed of input shaft IP1is reduced to a point identified by 2ND in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the second speed is shown inFIG. 10B. The power flows through first clutch C1, third brake B3, and the rotating members, as shown by bold lines, first planetary gearset G1and speed-shifting planetary gearset G21except second sun gear S2, as shown by a hatch pattern. In this speed, first planetary gearset G1and speed-shifting planetary gearset G21serve for the transmission of power and torque.

The third speed is established by releasing third brake B3and applying second brake B2to the operational state of the second speed, that is, by engaging first clutch C1and applying second brake B2, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. At speed-shifting planetary gearset G21, with first clutch C1engaged, the reduced speed is input to third sun gear S3from first planet-pinion carrier PC1of first planetary gearset G1. On the other hand, with second brake B2applied, second sun gear S2is fixed to the transmission housing. In this state, second sun gear S2, second planet-pinion carrier PC2that carries third planet pinion Pb21, and second ring gear R2are assumed to form a planetary gearset of the single-pinion type with a sun gear fixed. With second sun gear S2fixed and second ring gear R2rotating, second planet-pinion carrier PC2is forced to rotate at a speed reduced from the speed of second ring gear R2. At the same time, third sun gear S3, second planet-pinion carrier PC2that carries second planet pinion Pb1and third planet pinion Pb21, and second ring gear R2are assumed to form a planetary gearset of the double-pinion type with a planet-pinion carrier rotating at a low speed. Therefore, with third sun gear S3rotating at a speed reduced at first planetary gearset G1, second ring gear R2is forced to rotate at a further reduced speed (faster than the second speed), thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 9provides another analysis of the third speed. The point identified by C1in the diagram, or the engagement point of first clutch C1indicates the engagement of first clutch C1with which the reduced speed is input to third sun gear S3from first planetary gearset G1. The point identified by B2in the diagram, or the application point of second brake B2indicates the application of second brake B2with which second sun gear S2is held stationary. The operation of speed-shifting planetary gearset G21is defined by the lever or straight line connecting the engagement point of first clutch C1and the application point of second brake B2. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the third speed, the rotation speed of input shaft IP1is reduced to a point identified by 3RD in the diagram (faster than the second speed) through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the second speed is shown inFIG. 10C. The power flows through first clutch C1, second brake B2, and the rotating members, as shown by bold lines, and first planetary gearset G1and speed-shifting planetary gearset G21except fourth planet pinion Pb3and third ring gear R3, as shown by a hatch pattern.

The fourth speed is established by releasing second brake B2and engaging second clutch C2to the operational state of the third speed, that is, by engaging first clutch C1and second clutch C2, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. At speed-shifting planetary gearset G21, with first clutch C1engaged, the reduced speed is input to third sun gear S3from first planet-pinion carrier PC1of first planetary gearset G1. At the same time, with second clutch C2engaged, the reduced speed is input to second sun gear S2from first planet-pinion carrier PC1of first planetary gearset G1. Accordingly, at speed-shifting planetary gearset G21, third sun gear S3and second sun gear S2rotate at the same reduced speed, so that second planet-pinion carrier PC2and second ring gear R2also rotate solidly with third sun gear S3and second sun gear S2. Therefore second ring gear R2is forced to rotate at the reduced speed that is reduced at first planetary gearset G1, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 9provides another analysis of the fourth speed. The point identified by C1in the diagram, or the engagement point of first clutch C1indicates the engagement of first clutch C1with which the reduced speed is input to third sun gear S3from first planetary gearset G1. Similarly, the point identified by C2in the diagram, or the engagement point of second clutch C2indicates the engagement of second clutch C2with which the reduced speed is input to second sun gear S2from first planetary gearset G1. The operation of speed-shifting planetary gearset G21is defined by the lever or straight line connecting the engagement point of first clutch C1and the engagement point of second clutch C2. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the fourth speed, the rotation speed of input shaft IP1is reduced to a point identified by 4TH in the diagram (to the gear ratio of first planetary gearset G1) through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the second speed is shown inFIG. 11A. The power flows through first clutch C1, second clutch C2, and the rotating members, as shown by bold lines, and first planetary gearset G1and speed-shifting planetary gearset G21except fourth planet pinion Pb3and third ring gear R3, as shown by a hatch pattern.

The fifth speed is established by disengaging second clutch C2and engaging third clutch C3to the operational state of the fourth speed, that is, by engaging first clutch C1and third clutch C3, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. At speed-shifting planetary gearset G21, with first clutch C1engaged, the reduced speed is input to third sun gear S3from first planet-pinion carrier PC1of first planetary gearset G1. On the other hand, with third clutch C3engaged, the input rotation of input shaft IP1is input to second planet-pinion carrier PC2. Accordingly, in this state, third sun gear S3, second planet-pinion carrier PC2that carries second planet pinion Pb1and third planet pinion Pb21, and second ring gear R2are assumed to form a planetary gearset of the double-pinion type. Therefore, with third sun gear S3rotating at a speed reduced at first planetary gearset G1and second planet-pinion carrier PC2rotating at the input speed, second ring gear R2is forced to rotate at a speed intermediate between that of third sun gear S3and that of second planet-pinion carrier PC2, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 9provides another analysis of the fifth speed. The point identified by C1in the diagram, or the engagement point of first clutch C1indicates the engagement of first clutch C1with which the reduced speed is input to third sun gear S3from first planetary gearset G1. Similarly, the point identified by C3in the diagram, or the engagement point of third clutch C3indicates the engagement of third clutch C3with which the input speed is input to second planet-pinion carrier PC2from input shaft IP1. The operation of speed-shifting planetary gearset G21is defined by the lever or straight line connecting the engagement point of first clutch C1and the engagement point of third clutch C3. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the fifth speed, the rotation speed of input shaft IP1is reduced slightly to a point identified by 5TH in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the fifth speed is shown inFIG. 11B. The power flows through first clutch C1, third clutch C3, and the rotating members, as shown by bold lines, and first planetary gearset G1and speed-shifting planetary gearset G21except second sun gear S2, fourth planet pinion Pb3and third ring gear R3, as shown by a hatch pattern.

The sixth speed is established by disengaging first clutch C1and engaging second clutch C2to the operational state of the fifth speed, that is, by engaging second clutch C2and third clutch C3, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. With second clutch C2engaged, the reduced speed is input to second sun gear S2from first planet-pinion carrier PC1of first planetary gearset G1. At the same time, with third clutch C3engaged, the input rotation of input shaft IP1is input to second planet-pinion carrier PC2. In this state, second sun gear S2, second planet-pinion carrier PC2that carries third planet pinion Pb21, and second ring gear R2are assumed to form a planetary gearset of the single-pinion type. Therefore, with second sun gear S2rotating at a speed reduced at first planetary gearset G1and second planet-pinion carrier PC2rotating at the input speed, second ring gear R2is forced to rotate at a speed increased from the input speed, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 9provides another analysis of the sixth speed. The point identified by C2in the diagram, or the engagement point of second clutch C2indicates the engagement of second clutch C2with which the reduced speed is input to second sun gear S2from first planetary gearset G1. Similarly, the point identified by C3in the diagram, or the engagement point of third clutch C3indicates the engagement of third clutch C3with which the input speed is input to second planet-pinion carrier PC2from input shaft IP1. The operation of speed-shifting planetary gearset G21is defined by the lever or straight line connecting the engagement point of second clutch C2and the engagement point of third clutch C3. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the sixth speed, the rotation speed of input shaft IP1is increased slightly to a point identified by 6TH in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the sixth speed is shown inFIG. 1C. The power flows through second clutch C2, third clutch C3, and the rotating members, as shown by bold lines, and first planetary gearset G1and speed-shifting planetary gearset G21except third sun gear S3, second planet pinion Pb1, fourth planet pinion Pb3, and third ring gear R3, as shown by a hatch pattern.

The seventh speed is established by disengaging second clutch C2and applying second brake B2to the operational state of the sixth speed, that is, by engaging third clutch C3and applying second brake B2, as shown inFIG. 2. With third clutch C3engaged, the input rotation speed of input shaft IP1is input to second planet-pinion carrier PC2of speed-shifting planetary gearset G21. On the other hand, with second brake B2applied, second sun gear S2of speed-shifting planetary gearset G21is held stationary to the transmission housing. In this state, second sun gear S2, second planet-pinion carrier PC2that carries third planet pinion Pb21, and second ring gear R2are assumed to form a planetary gearset of the single-pinion type with a sun gear fixed. Therefore, with second planet-pinion carrier PC2rotating at the input speed, second ring gear R2is forced to rotate at a speed increased faster than the input speed, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 9provides another analysis of the seventh speed. The point identified by C3in the diagram, or the engagement point of third clutch C3indicates the engagement of third clutch C3with which the input speed is input to second planet-pinion carrier PC2from input shaft IP1. The point identified by B2in the diagram, or the application point of second brake B2indicates the application of second brake B2with which second sun gear S2is held stationary. The operation of speed-shifting planetary gearset G21is defined by the lever or straight line connecting the engagement point of third clutch C3and the application point of second brake B2. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the seventh speed, the rotation speed of input shaft IP1is increased to a point identified by 7TH in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the seventh speed is shown inFIG. 12A. The power flows through third clutch C3, second brake B2, and the rotating members, as shown by bold lines, and speed-shifting planetary gearset G21except third sun gear S3, second planet pinion Pb1, fourth planet pinion Pb3, and third ring gear R3, as shown by a hatch pattern.

The eighth speed is established by releasing second brake B2and applying third brake B3to the operational state of the seventh speed, that is, by engaging third clutch C3and applying third brake B3, as shown inFIG. 2. With third clutch C3engaged, the input rotation speed of input shaft IP1is input to second planet-pinion carrier PC2of speed-shifting planetary gearset G21. On the other hand, with third brake B3applied, third ring gear R3of speed-shifting planetary gearset G21is held stationary to the transmission housing. The rotation of third planet pinion Pb21is defined by the rotations of second planet-pinion carrier PC2and third ring gear R3via fourth planet pinion Pb3. Therefore, second ring gear R2is forced to rotate at a speed defined and increased by the rotations of second planet-pinion carrier PC2and third planet pinion Pb21.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 9provides another analysis of the eighth speed. The point identified by C3in the diagram, or the engagement point of third clutch C3indicates the engagement of third clutch C3with which the input speed is input to second planet-pinion carrier PC2from input shaft IP1. The point identified by B3in the diagram, or the application point of third brake B3indicates the application of third brake B3with which third ring gear R3is held stationary to the transmission housing. The operation of speed-shifting planetary gearset G21is defined by the lever or straight line connecting the engagement point of third clutch C3and the application point of third brake B3. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the eighth speed, the rotation speed of input shaft IP1is increased to a point identified by 8TH in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the eighth speed is shown inFIG. 12B. The power flows through third clutch C3, third brake B3, and the rotating members, as shown by bold lines, and speed-shifting planetary gearset G21except second sun gear S2, third sun gear S3, and second planet pinion Pb1, as shown by a hatch pattern.

The first reverse speed is established by engaging second clutch C2and applying first brake B1, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. With second clutch C2engaged, the reduced speed is input to second sun gear S2from first planet-pinion carrier PC1of first planetary gearset G1. On the other hand, with first brake B1applied, second planet-pinion carrier PC2is fixed to the transmission housing. Second sun gear S2, second planet-pinion carrier PC2that carries third planet pinion Pb21, and second ring gear R2are assumed to form a planetary gearset of the single-pinion type with a planet-pinion carrier fixed. Therefore, with second sun gear S2rotating at the reduced speed, second ring gear R2is forced to rotate in the reverse direction at a reduced speed, thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 9provides another analysis of the first reverse speed. The point identified by C2in the diagram, or the engagement point of second clutch C2indicates the engagement of second clutch C2with which the reduced speed is input to second sun gear S2from first planetary gearset G1. The point identified by B1in the diagram, or the application point of first brake B1indicates the application of first brake B1with which second planet-pinion carrier PC2is held stationary. The operation of speed-shifting planetary gearset G21is defined by the lever or straight line connecting the engagement point of second clutch C2and the application point of first brake B1. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the first reverse speed, the rotation speed of input shaft IP1is reduced to a point identified by REV1 in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the first reverse speed is shown inFIG. 13A. The power flows through second clutch C2, first brake B1, and the rotating members, as shown by bold lines, and first planetary gearset G1and speed-shifting planetary gearset G21except third sun gear S3, second planet pinion Pb1, fourth planet pinion Pb3, and third ring gear R3, as shown by a hatch pattern.

The second reverse speed is established by releasing first brake B1and applying third brake B3to the operational state of the first reverse speed, that is, by engaging second clutch C2and applying third brake B3, as shown inFIG. 2. First, the rotation speed of input shaft IP1is reduced at first planetary gearset G1. With second clutch C2engaged, the reduced speed is input to second sun gear S2from first planet-pinion carrier PC1of first planetary gearset G1. On the other hand, with third brake B3applied, third ring gear R3is fixed to the transmission housing. In this state, second sun gear S2, second planet-pinion carrier PC2that carries third planet pinion Pb21and fourth planet pinion Pb3, and third ring gear R3are assumed to form a planetary gearset of the double-pinion type with a ring gear fixed. With third ring gear R3fixed and second sun gear S2rotating at the reduced speed, second planet-pinion carrier PC2is forced to rotate in the reverse direction at a speed reduced from the speed of second sun gear S2. At the same time, second sun gear S2, second planet-pinion carrier PC2that carries third planet pinion Pb21, and second ring gear R2are assumed to form a planetary gearset of the single-pinion type with a planet-pinion carrier rotating at a low speed. Therefore, with second sun gear S2rotating at a speed reduced at first planetary gearset G1, second ring gear R2is forced to rotate in the reverse direction at a reduced speed (faster than the first reverse speed), thereby outputs the rotation speed to output shaft OP1.

The collinear diagram of the multiple-speed automatic transmission as shown inFIG. 9provides another analysis of the second reverse speed. The point identified by C2in the diagram, or the engagement point of second clutch C2indicates the engagement of second clutch C2with which the reduced speed is input to second sun gear S2from first planetary gearset G1. The point identified by B3in the diagram, or the application point of third brake B3indicates the application of third brake B3with which third ring gear R3is held stationary. The operation of speed-shifting planetary gearset G21is defined by the lever or straight line connecting the engagement point of second clutch C2and the application point of third brake B3. The intersection point of the lever and the perpendicular line at output shaft OP1indicates the output speed. In the second reverse speed, the rotation speed of input shaft IP1is reduced to a point identified by REV2 in the diagram through the multiple-speed automatic transmission, and output to output shaft OP1.

The power flow or torque flow in the second reverse speed is shown inFIG. 13B. The power flows through second clutch C2, third brake B3, and the rotating members, as shown by bold lines, and first planetary gearset G1and speed-shifting planetary gearset G21except third sun gear S3and second planet pinion Pb1, as shown by a hatch pattern.

The multiple-speed automatic transmission of the second embodiment is constructed and operated as discussed above. The following describes effects of the second embodiment. The multiple-speed automatic transmission of the second embodiment provides the same effects (E1), (E2), and (E3) as the first embodiment, and additionally the following effects. (E4) Employing a pinion gear with a single diameter or a single number of teeth as third planet pinion Pb21results in simple and easy processing.

(E5) Placing fourth planet pinion Pb3and third ring gear R3outside second sun gear S2results in multiplication of gear ratios without extending the overall longitudinal length.

The gearbox of the present invention, which is applied to the multiple-speed automatic transmission in the embodiments, may be applied to other types of transmissions.

The entire contents of Japanese Patent Application No. 2003-202222 (filed Jul. 28, 2003) are incorporated herein by reference.

While the foregoing is a description of the preferred embodiments carried out the invention, it will be understood that the invention is not limited to the particular embodiments shown and described herein, but that various changes and modifications may be made without departing from the scope or spirit of this invention as defined by the following claims.