Multi speed power transmission

A powertrain has a multi speed power transmission including a planetary gear arrangement with three simple planetary gear sets and six selectively engageable, fluid operated, friction torque transmitting mechanisms. A one-way torque transmitting mechanism is disposed in parallel drive relation with one of the friction torque transmitting mechanisms. The selective engagement of the friction torque transmitting mechanisms will establish seven forward drive ratios, a neutral condition and a reverse drive ratio. All of the forward ratio interchanges are single transition shifts and all of the single ratio skip shifts are single transition shifts. The seven forward drive ratios include four underdrive ratios, a direct drive ratio, and two overdrive ratios. The planetary gear arrangement can be controlled to establish six forward drive ratios with either one overdrive ratio and four underdrive ratios or two overdrive ratios and three underdrive ratios. A five speed arrangement having four forward underdrive ratios, a forward direct ratio a neutral condition and a reverse drive ratio while permitting the elimination of one of the friction torque transmitting mechanisms.

TECHNICAL FIELD
 This invention relates to power transmissions and more particularly to
 multi speed power transmissions having three simple planetary gear sets.
 BACKGROUND OF THE INVENTION
 Many of the current automobile manufacturers are using five speed automatic
 transmissions that incorporate three planetary gear sets. There has also
 been a number of recently issued patents issued that describe five speed
 automatic transmissions having three planetary gear sets; and some
 describing five speed automatic transmissions having two planetary gear
 sets that include a continuous drive connection and a selective drive
 connection between the planetary gear sets. Most of the two planetary gear
 set transmissions incorporate a simple planetary gear set and a compound
 planetary gear set. The transmissions having three simple planetary gear
 sets have at least one continuous drive connection uniting a member of
 each planetary gear set; or a clutch mechanism that will unite the three
 gear sets. These multi speed power transmissions include at least one
 selectively engageable torque transmitting mechanism for each of the
 forward speeds. In other words, a five speed transmission has five
 selectively engageable torque transmitting mechanisms. Generally, the
 torque transmitting mechanisms include three clutches and two brakes
 A transmission described in U.S. Pat. No. 4,070,927, issued to Polak on
 Jan. 31, 1978, has three simple planetary gear sets controlled by five
 torque transmitting mechanisms to produce six forward speed ratios and a
 reverse ratio. This transmission produces three underdrive ratios, a
 direct drive ratio and two overdrive ratios in the forward direction. To
 provide a seventh speed, Polak incorporates an additional planetary gear
 set.
 SUMMARY OF THE INVENTION
 It is an object of the present invention to provide an improved multi speed
 power transmission having three simple planetary gear sets.
 In one aspect of the invention, the three simple planetary gear sets are
 controlled by a plurality of selectively engageable torque transmitting
 mechanisms to provide seven forward speed ratios, a reverse ratio, and a
 neutral condition. In another aspect of the present invention, the seven
 forward speed ratios, with coast braking, and the reverse ratio are
 established through the judicious engagement of six torque transmitting
 mechanisms.
 In yet another aspect of the present invention, a one-way torque
 transmitting device provides the reaction grounding mechanism for the
 first and second forward ratios when coast braking is not desired. In
 still another aspect of the present invention, two of the planetary gear
 sets each have a member connected for continuous rotation with an output
 shaft and each have another member connected for common rotation.
 In yet still another aspect of the present invention, each of the planetary
 gear sets has at least one member connectable with an input shaft through
 a selectively engageable clutch mechanism. In a further aspect of the
 present invention, the third of the planetary gear sets has a member
 connected for common rotation with a member of one of the other two
 planetary gear sets and with a selectively engageable friction brake. In
 yet a further aspect of the present invention, the three planetary gear
 sets are controlled by the selectively engageable torque transmitting
 mechanisms to produce four underdrive ratios, a direct drive ratio and two
 overdrive ratios.
 In still a further aspect of the present invention, the three planetary
 gear sets can be controlled to provide a six speed transmission having
 four underdrive ratios and one overdrive ratio, a six speed transmission
 having three underdrive ratios and two overdrive ratios, and a five speed
 transmission having four underdrive ratios and a direct drive ratios with
 only two selectively friction engageable clutches and three selectively
 engageable friction brakes. In a yet still further aspect of the present
 invention, one of the planetary gear sets has one member continuously
 drive connected with the input shaft as well as one member selectively
 connectable with the input shaft through a friction clutch.

DESCRIPTION OF AN EXEMPLARY EMBODIMENT
 A powertrain 10, shown in FIG. 1, has an engine 12, a torque converter 14,
 a multi speed transmission 16 and a final drive 18. The engine 12 is a
 conventional prime mover such as an internal combustion engine. The torque
 converter is a conventional hydrodynamic device and the final drive is a
 conventional reduction and differential gear mechanism
 The transmission 16 has a planetary gear arrangement including three simple
 planetary gear sets 20, 22 and 24; three conventional selectively
 engageable, fluid operated, rotating torque transmitting mechanisms 26, 28
 and 30; three conventional selectively engageable, fluid operated,
 stationary torque transmitting mechanisms 32, 34, and 36 and a one-way
 torque transmitting mechanism 38. The one-way torque transmitting
 mechanism 38 and the torque transmitting mechanism 36 are disposed in
 parallel torque transmitting paths to a transmission housing 42.
 The torque transmitting mechanisms 26, 28 and 30 are drivingly connected
 with an input housing 44 that is drivingly connected with the torque
 converter 14 through an input shaft 46. The housing 44 is connected via a
 hub 48 with a sun gear member 50 that is a component of the planetary gear
 set 20. The planetary gear set 20 also includes a ring gear member 52 and
 a carrier assembly member 54. The carrier assembly member 54 has a
 plurality of rotatably mounted pinion gear members 56 that mesh with both
 the sun gear member 50 and the ring gear member 52. The carrier assembly
 member 54 is selectively connectable with the input housing 44 through the
 torque transmitting mechanism 30 and with the transmission housing 42
 through the torque transmitting mechanism 32. The ring gear member 52 is
 selectively connectable with the housing 42 through the torque
 transmitting mechanism 34 and continuously connected with a sun gear
 member 58 which is a member of the planetary gear set 22.
 The planetary gear set 22 also includes a ring gear member 60 and a carrier
 assembly member 62 that has a plurality of rotatably mounted pinion gear
 members 64 disposed in meshing relation with both the sun gear member 58
 and the ring gear member 60. The carrier assembly member 62 is selectively
 connectable with the input housing 44 through the torque transmitting
 mechanism 28 and is also selectively connectable with the transmission
 housing 42 through the torque transmitting mechanism 36 and the one-way
 torque transmitting mechanism 38. The carrier assembly member 62 is
 continuously drive connected with a ring gear member 66 that is a
 component of the planetary gear set 24.
 The planetary gear set 24 also includes a sun gear member 68 and a carrier
 assembly member 70 which has a plurality of rotatably mounted pinion gear
 members 72 disposed in meshing relation with the sun gear member 68 and
 the ring gear member 66. The sun gear member 68 is selectively connectable
 with the input housing 44 through the torque transmitting mechanism 26.
 The carrier assembly member 70 and the ring member 60 of the planetary
 gear set 22 are continuously drivingly connected with an output shaft 74
 that is drivingly connected with the final drive mechanism 18.
 The planetary gear arrangement 16 will provide a reverse drive ratio, a
 neutral condition and seven forward drive ratios between the input shaft
 46 and the output shaft 74 when the torque transmitting mechanisms 26, 28,
 20, 32, 34, and 36 are selectively engaged in accordance with the scheme
 set forth in the truth table shown in FIG. 2. The engagement and
 disengagement of the torque transmitting mechanisms 26, 28, 20, 32, 34,
 and 36 is preferably controlled by a conventional electronic control unit
 (ECU), not shown, that includes a preprogrammed digital computer in a
 well-known manner.
 To establish the reverse drive ratio in the planetary gear arrangement 16,
 the torque transmitting mechanisms 36 and 30 are engaged. As indicated in
 FIG. 2, the torque transmitting mechanism 30 is engaged to initiate
 vehicle launch in reverse. To accomplish this the engagement of torque
 transmitting mechanism 30 occurs at a rate controlled by the ECU in a
 well-known manner. With the engagement of the torque transmitting
 mechanism 36, the carrier assembly member 62 is stationary and the sun
 gear member 58 is driven forwardly, engine rotation direction, such that
 the ring gear member 60 and output shaft 74 rotate in reverse.
 The neutral condition of the planetary gear arrangement 16 is established
 when all of the torque transmitting mechanisms are disengaged. However,
 the torque transmitting mechanism 36 can be active without a drive ratio
 being established. This will permit a shift from reverse to first forward
 or from neutral to either reverse or first forward with only the garage
 shift element being controlled.
 To establish the first forward ratio, the torque transmitting mechanism 32
 is engaged at a controlled rate. Either the torque transmitting mechanism
 36 or the one-way torque transmitting mechanism 38 will establish the
 carrier assembly member 62 as a reaction member. The controlled engagement
 of the torque transmitting mechanism 32 will establish the carrier
 assembly member 54 as a reaction member. The sun gear member 50 will
 provide an input member for the planetary gear arrangement 16. The
 planetary gear sets 20 and 22 provide the reduction ration for the first
 forward drive ratio.
 To establish the second forward drive ratio, the torque transmitting
 mechanisms 32 and 26 are interchanged under the control of the ECU while
 the torque transmitting mechanism 38 remains engaged. The sun gear member
 68 becomes the input member and the ring gear member 66 becomes the
 reaction member. The ring gear member 66 is restrained from rotation by
 the torque transmitting mechanism 36, if engaged, or the one-way torque
 transmitting mechanism 38. This is a single transition shift in that only
 one torque transmitting mechanism needs to be released and only one torque
 transmitting mechanism need to be engaged. The release and engagement of
 the torque transmitting mechanisms is controlled by the ECU in a
 well-known manner. The second ratio is provided through the planetary gear
 set 24.
 To establish the third forward ratio, the one-way torque transmitting
 mechanism 38 and the torque transmitting mechanism 32 are interchanged, as
 controlled by the ECU, while the torque transmitting mechanism 26 remains
 engaged. This is a single transition shift. The one-way torque
 transmitting mechanism 38 will automatically release when the ring gear
 member 66 rotates forwardly. The third forward ratio is an underdrive
 ratio established by all three planetary gear sets 20, 22 and 24.
 The fourth forward drive ratio is established by the interchange of the
 torque transmitting mechanisms 32 and 34, as controlled by the ECU, while
 the torque transmitting mechanism 26 remains engaged. This is a single
 transition shift. The engagement of the torque transmitting mechanism 34
 will establish the sun gear member 58 as a reaction member in the
 planetary gear arrangement 16. The sun gear member 68 remains the input
 member. The fourth forward ratio is an underdrive ratio established by the
 planetary gear sets 22 and 24.
 The fifth forward drive ratio is established by the interchange of the
 torque transmitting mechanisms 34 and 28, as controlled by the ECU, while
 the torque transmitting mechanism 26 remains engaged. This is a single
 transition shift. In the fifth forward drive ratio two input torque
 transmitting mechanisms are engaged which will result in the planetary
 gear arrangement rotating as a single unit to provide a direct drive. It
 should be noted that both the ring gear member 66 and the sun gear member
 68 are rotated at the speed of the input shaft 46 such that the carrier
 assembly member 70 and the output shaft 74 will also rotate at the speed
 of the input shaft 46.
 The sixth forward drive ratio is established by the interchange of the
 torque transmitting mechanisms 26 and 34 under the control of the ECU
 while torque transmitting mechanism 28 remains engaged. This is a single
 transition shift. The torque transmitting mechanism 34 establishes the sun
 gear member 58 as the reaction member in the planetary gear arrangement 16
 and the torque transmitting mechanism 28 establishes the input member in
 the planetary gear arrangement 16. This presents an overdrive ratio
 between the input shaft 46 and the output shaft 74 through the planetary
 gear set 22.
 The seventh forward drive ratio is established by the interchange of the
 torque transmitting mechanisms 34 and 32 under the control of the ECU
 while the torque transmitting mechanism 28 remains engaged. The carrier
 assembly member 62 and the sun gear member 50 are the input members for
 the planetary gear arrangement 16 and carrier assembly member 54 becomes
 the reaction member for the planetary gear arrangement 16. The planetary
 gear sets 20 and 22 establish the seventh forward ratio, which is an
 overdrive ratio between the input shaft 46 and the output shaft 74.
 While the sun gear member 50 is continually driven by the input shaft 46,
 the planetary gear set 20 only contributes to the drive ratio when the
 torque transmitting mechanism 32 is engaged. This occurs in the first,
 third, and seventh forward ratios.
 The truth table depicted in FIG. 2 presents a typical set of drive ratios
 that are available with the present invention. The "OAR" term in the truth
 table is the overall ratio between the first forward ratio and the seventh
 forward ratio. The step column represents the step size between adjacent
 ratios, for example, the step size between the first and second forward
 ratios is 1.81. The Ri/Si represents the ratio of the number of teeth on
 the ring gear 52 to the number of teeth on the sun gear 50. The R2/S2 and
 R3/S3 represent corresponding values for the respective gears in the
 planetary gear sets 22 and 24.
 The truth table in FIG. 2 also makes it apparent that the single ratio skip
 shifts (i.e. first to third, second to fourth etc.) are single transition
 shifts. Only one torque transmitting mechanism is disengaged while only
 one torque transmitting mechanism is engaged. During a fourth to sixth
 ratio interchange, the torque transmitting device 26 is disengaged while
 the torque transmitting device 28 is engaged. Those skilled in the art
 will recognize the other single transition skip shifts. The overall ratio
 of 7.5 and acceptable ratio step provide an excellent transmission for
 truck applications. The large first gear ratio (4.870) provides very good
 towing capacity and enables the torque converter to be downsized. The
 large first gear ratio and the close (small step) ratios will allow for
 engine down sizing in some applications. The second forward drive ratio is
 sufficiently high to permit second gear launch when the vehicle is not
 heavily loaded. During a second ratio start, the torque transmitting
 mechanism 26 would be engaged to control vehicle launch.
 Two six speed transmission are available with the present invention. One of
 the six speed transmissions drops the seventh ratio to provide four
 underdrives, a direct drive and one overdrive. The reverse drive is
 unchanged. The low or first gear ratio remains a 4.870 to provide good
 launch characteristics and towing for both a truck or automobile
 transmission. The sixth gear ratio become 0.742 and the OAR is 6.6.
 The other six speed drops the low gear ratio and changes the R1/S1 and
 R3/S3 ratios to 1.96 by changing the number of teeth on the sun gears 50
 and 68 from 58 teeth to 50 teeth. This establishes the first gear ratio at
 2.96 and the sixth gear ratio at 0.656 with the OAR at 4.5. This will
 provide a transmission that is suited for applications in rear wheel drive
 automobiles. The step ratios are close to establish very good acceleration
 performance and ratio interchange smoothness. Those skilled in the art
 will recognize that the low gear ratio of this proposal is determined by
 the planetary gear set 24 and the second and sixth forward ratios are
 partly established by the planetary gear set 20. The ratio interchange
 pattern will be the same as that shown in FIG. 2 for the second through
 seventh ratios.
 One other variation is possible with the present invention; and that is a
 power transmission with five forward ratios, a neutral condition and one
 reverse ratio. This transmission presents four underdrive ratios and a
 direct drive ratio in the forward direction. Other commonly used five
 speed transmissions have at least one overdrive ratio and most have two
 overdrive ratios. The torque transmitting mechanism 28 is eliminated when
 a five speed transmission is desired. This will shorten the overall length
 of the transmission for improved packaging and space utilization,
 particularly in front wheel drive applications. The five speed
 transmission would preferably use the ratios set forth for first through
 fifth in FIG. 2. An OAR of 4.9 is available with these values and the
 launch ratio of 4.87 is retained.
 With the elimination of the torque transmitting mechanism 28, the fifth
 ratio is established with the engagement of the torque transmitting
 mechanisms 26 and 30. Thus the torque transmitting mechanism 30 is
 employed in both the reverse ratio and the fifth forward ratio. The
 interchange between fourth and fifth is accomplished through the
 interchange of the torque transmitting mechanisms 34 and 30. This is a
 single transition shift and the skip shifts remain single transition
 shifts. The large launch ratio (4.87) permits downsizing of the torque
 converter and the direct drive (one to one) fifth ratio allows the use of
 low transfer gear ratios in the final drive 18 which provides efficient
 top gear fuel economy.
 Those skilled in the art will recognize that the rotating torque
 transmitting mechanisms 26, 28 and 30 are commonly designated as clutches
 and the stationary torque transmitting mechanisms 32, 34 and 36 are
 commonly termed brakes or stationary clutches.