Patent Abstract:
A six speed transmission is provided that includes three planetary gear sets having a common carrier member and six torque-transmitting mechanisms operated in combinations of two to provide at least six forward speed ratios and two reverse speed ratios. A method of assembling a transmission is also provided. A reduction in components and component standardization is achieved.

Full Description:
TECHNICAL FIELD 
     The present invention relates to power transmissions having three planetary gear sets with a single, common carrier member. More specifically, the gear sets are controlled by six torque-transmitting mechanisms to provide six forward speed ratios and two reverse speed ratios. 
     BACKGROUND OF THE INVENTION 
     Passenger vehicles include a powertrain that is comprised of an engine, multi-speed transmission, and a differential or final drive. The multi-speed transmission increases the overall operating range of the vehicle by permitting the engine to operate through its torque range a number of times. The number of forward speed ratios that are available in the transmission determines the number of times the engine torque range is repeated. Early automatic transmissions had two speed ranges. This severely limited the overall speed range of the vehicle and therefore required a relatively large engine that could produce a wide speed and torque range. This resulted in the engine operating at a specific fuel consumption point during cruising, other than the most efficient point. Therefore, manually-shifted (countershaft transmissions) were the most popular. 
     With the advent of three- and four-speed automatic transmissions, the automatic shifting (planetary gear) transmission increased in popularity with the motoring public. These transmissions improved the operating performance and fuel economy of the vehicle. The increased number of speed ratios reduces the step size between ratios and therefore improves the shift quality of the transmission by making the ratio interchanges substantially imperceptible to the operator under normal vehicle acceleration. 
     It has been suggested that the number of forward speed ratios be increased to six or more. Six-speed transmissions are disclosed in U.S. Pat. No. 4,070,927 issued to Polak on Jan. 31, 1978; and U.S. Pat. No. 6,422,969 issued to Raghavan and Usoro on Jul. 23, 2002. 
     Six-speed transmissions offer several advantages over four- and five-speed transmissions, including improved vehicle acceleration and improved fuel economy. While many trucks employ power transmissions having six or more forward speed ratios, passenger cars are still manufactured with three- and four-speed automatic transmissions and relatively few five or six-speed devices due to the size and complexity of these transmissions. 
     SUMMARY OF THE INVENTION 
     A multi-speed transmission is provided having a single, common carrier member functioning for each of multiple planetary gear sets. The single carrier member allows for reduction in components and a potentially lower transmission cost. 
     Accordingly, a multi-speed transmission includes an input shaft and an output shaft. First, second and third planetary gear sets have a common carrier member, at least two ring gear members, and each has a sun gear member. A plurality of sets of pinion gears is rotatably mounted on the common carrier member for intermeshing with the ring gear members and the sun gear members. The input shaft is not continuously connected with any member of the planetary gear sets and the output shaft is continuously connected with a member of the planetary gear sets. Six torque-transmitting mechanisms are operable for selectively interconnecting members of the planetary gear sets with the input shaft, with a stationary member, or with other members of the planetary gear sets. The six torque-transmitting mechanisms are engaged in combinations of two to establish at least six forward speed ratios and two reverse speed ratios between the input shaft and the output shaft. 
     In one aspect of the invention, an interconnecting member continuously interconnects the sun gear member of one of the planetary gear sets with the sun gear member of another of the planetary gear sets. 
     In another aspect of the invention, one of the ring gear members intermeshes with pinion gears of both the first and second planetary gear sets. 
     In yet another aspect of the invention, each pinion gear is characterized by a predetermined number of teeth and rotates on a respective spindle mounted on the common carrier member at a respective bearing. Each of the spindles is of the same size and each of the bearings is of the same size. 
     The first and second planetary gear sets may have separate ring gear members or share a common ring gear member. If separate ring gear members are used for the first and second planetary gear sets, an interconnecting member interconnects the ring gear member of the first planetary gear set with the ring gear member of the second planetary gear set. 
     In yet another aspect of the invention, a first of the six torque-transmitting mechanisms is operable for selectively interconnecting the sun gear member of the second planetary gear set and the sun gear member of the third planetary gear set with the input shaft. 
     In still another aspect of the invention, a second of the six torque-transmitting mechanisms is operable for selectively interconnecting the common carrier member with the input shaft. 
     In another aspect of the invention, a third of the six torque-transmitting mechanisms is operable for selectively interconnecting the input shaft with the sun gear member of the first planetary gear set. Preferably the third torque-transmitting mechanism is disposed radially inward of the first and second torque-transmitting mechanisms, as the third torque-transmitting mechanism is characterized by higher speeds than the first and second torque-transmitting mechanisms. The radially inward position of the third torque-transmitting mechanism thereby minimizes spin losses. 
     In still another aspect of the invention, a fourth of the six torque-transmitting mechanisms is operable for selectively interconnecting the sun gear member of the first planetary gear set with the stationary member. 
     In still further aspect of the invention, a fifth of the six torque-transmitting mechanisms is operable for selectively interconnecting a ring gear member of the first and second planetary gear sets with the stationary member. 
     In another aspect of the invention, a sixth of the six torque-transmitting mechanisms is operable for selectively interconnecting the common carrier member with the stationary member. 
     A method of assembling a transmission having multiple planetary gear sets includes providing a single carrier member configured to rotatably support pinion gears for each of the planetary gear sets. The method also includes radially positioning the single carrier member within the transmission on two axially spaced bushings. Thus, the use of the single carrier member enables relatively simple and accurate positioning during assembly. 
     The above features and advantages and other features and advantages of the present invention are readily apparent from the following detailed description of the best modes for carrying out the invention when taken in connection with the accompanying drawings. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a schematic representation of a powertrain including an embodiment of a planetary transmission of the present invention with an alternative single ring gear member shown in phantom; 
         FIG. 2  is a truth table depicting some of the operating characteristics of the powertrain shown in  FIG. 1 ; and 
         FIG. 3  is a chart depicting other operating characteristics of the powertrain shown in  FIG. 1 . 
     
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Referring to the drawings, wherein like reference numerals represent the same or corresponding parts through the several views, there is shown in  FIG. 1  a powertrain  10  having a conventional engine and torque converter  12 , a planetary transmission  14  and a conventional final drive mechanism  16 . 
     The planetary transmission  14  includes an input shaft  17  continuously connected with the engine and torque converter  12 , a planetary gear arrangement  18 , and an output shaft  19  continuously connected with the final drive mechanism  16 . The planetary gear arrangement  18  includes three planetary gear sets  20 ,  30  and  40 . 
     The planetary gear set  20  includes a sun gear member  22 , a ring gear member  24 , and a planet carrier assembly member  26 . The planet carrier assembly member  26  includes a plurality of pinion gears  27  (a first set of pinion gears) rotatably mounted on a carrier member  29  and disposed in meshing relationship with the sun gear member  22 . A plurality of pinion gears  28  (a second set of pinion gears) is also rotatably mounted on the carrier member  29 . The pinion gears  28  are disposed in meshing relationship with the pinion gears  27  and the ring gear member  24 . 
     The planetary gear set  30  includes a sun gear member  32 , a ring gear member  34 , and the same planet carrier assembly member  26 . The planet carrier assembly member  26  includes a plurality of pinion gears  37  rotatably mounted on the same carrier member  29  and disposed in meshing relationship with both the sun gear member  32  and the ring gear member  34 . Pinion gears  37  are also referred to herein as a third set of pinion gears. 
     In an alternative embodiment, rather than separate ring gear members  24  and  34 , a single ring gear member  24 ′ is included in both the planetary gear sets  20  and  30 . The single ring gear member  24 ′ is in meshing relationship with both the pinion gears  28  and the pinion gears  37 . 
     The planetary gear set  40  includes a sun gear member  42 , a ring gear member  44 , and the same planet carrier assembly member  26 . The planet carrier assembly member  26  includes a plurality of pinion gears  47  (a fourth set of pinion gears) rotatably mounted on the same carrier member  29  and disposed in meshing relationship with the sun gear member  42 . A plurality of pinion gears  48  (a fifth set of pinion gears) is also rotatably mounted on the carrier member  29  and is disposed in meshing relationship with both the ring gear member  44  and the pinion gears  47 . 
     Each pinion gear  27 ,  28 ,  37 ,  47 , and  48  rotates on a spindle mounted on the common carrier member  29  at a respective bearing. For instance, pinion gear  27  rotates on a spindle  25 A mounted on carrier member  29  at bearing  21 A. Pinion gear  28  rotates on a spindle  25 B mounted on carrier member  29  at bearing  21 B. Pinion gear  37  rotates on spindle  25 C mounted on carrier member  29  at bearing  21 C. Pinion gear  47  rotates on spindle  25 D mounted on carrier member  29  at bearing  21 D. Pinion gear  48  rotates on spindle  25 E mounted on common carrier member  29  at bearing  21 E. 
     The use of the common carrier member  29  reduces the number of required housing walls, saving axial space and reducing the number of thrust washers required. For instance, thrust washers (not shown) may be positioned adjacent end housing walls  23 A and  23 D to absorb axial thrust. However, no thrust washers are required between gear sets  20  and  30  or between gear sets  30  and  40 , as shared housing walls  23 B,  23 C, respectively, are employed. 
     The input shaft  17  is not continuously connected with any member of the planetary gear sets  20 ,  30  and  40 . The output shaft  19  is continuously connected with the ring gear member  44  via a drum  97 . A park lock gear  80  is also disposed on the drum  97  such that it is continuously interconnected with the ring gear member  44  and the output member  19 . The sun gear member  32  is continuously connected with the sun gear member  42  via an interconnecting member  70 . The ring gear member  24  is continuously connected with the ring gear member  34  via a drum  96 . Those skilled in the art will readily understand that a consolidation of parts is realized by utilizing an alternative embodiment having a single ring gear member  24 ′ in lieu of the separate ring gear members  24  and  34 . 
     The input shaft  17  is selectively connectable with the sun gear member  32  via torque-transmitting mechanism  50  which may be referred to herein as the C 1  clutch. Selective engagement of the C 1  clutch connects a drum  92  which is continuously connected with the input member  17  to an inner shaft  95  which is continuously connected with the sun gear member  32 . The input shaft  17  is also selectively connectable with the common carrier member  29  via a torque-transmitting mechanism  52 , which may also be referred to herein as the C 2  clutch. The C 2  clutch  52  selectively interconnects the drum  92  with an intermediate shaft  94  that is continuously connected with the carrier member  29 . Additionally, the input shaft  17  is also selectively connectable with the sun gear member  22  via a torque-transmitting mechanism  54 , which may also be referred to herein as C 3  clutch  54 . The C 3  clutch  54  selectively interconnects the drum  92  with an outer shaft  93  that is continuously connected with the sun gear member  22 . Torque-transmitting mechanism  56  which may also be referred to herein as the C 4  clutch  56 , selectively connects the sun gear member  22  with the transmission housing  60 , which may also be referred to herein as a stationary member. An extension or wall  62  of the transmission housing  60  extends between the drum  92  and the C 4  clutch  56 . The wall  62  provides support and oil feed for the clutches  50 ,  52  and  54 . A torque-transmitting mechanism  58 , which may also be referred to herein as the C 5  clutch, selectively connects the drum  96  with the transmission housing  60 , thereby grounding the ring gear members  24  and  34  with the transmission housing  60 . If the alternative single ring gear member  24 ′ is used, the C 5  clutch  58  selectively interconnects the ring gear member  24 ′ with the transmission housing  60 . A torque-transmitting mechanism  59  selectively connects the common carrier member  29  with the transmission housing  60 . The torque-transmitting mechanism  59  may also be referred to herein as the C 6  clutch. 
     The Reverse  1  speed ratio is established with the engagement of the C 3  clutch  54  and the C 5  clutch  58 . The C 3  clutch  54  connects the input member  17  with the sun gear member  22 , and the C 5  clutch  58  connects the ring gear member  24  and the ring gear member  34  with the transmission housing  60 . The sun gear member  22  rotates at the same speed as the input shaft  17 . The ring gear member  24  and the ring gear member  34  do not rotate. If the alternative embodiment having the single ring gear member  24 ′ is utilized, the ring gear member  24 ′ does not rotate. The common carrier member  29  rotates as a speed determined from the speed of the sun gear member  22  and the ring gear/sun gear tooth ratio of the planetary gear set  20 . The sun gear member  32  rotates at the same speed as the sun gear member  42 . The sun gear member  32  rotates at a speed determined from the speed of the common carrier member  29  and the ring gear/sun gear tooth ratio of the planetary gear set  30 . The ring gear member  44  rotates at the same speed as the output shaft  19 . The ring gear member  44 , and therefore the output shaft  19 , rotates at a speed determined from the speed of the common carrier member  29 , the speed of the sun gear member  42 , and the ring gear/sun gear tooth ratio of the planetary gear set  40 . The numerical value of the Reverse  1  speed ratio is determined utilizing the ring gear/sun gear tooth ratios of the planetary gear sets  20 ,  30  and  40 . 
     The Reverse Low speed ratio is established with the engagement of the C 3  clutch  54  and the C 6  clutch  59 . The C 3  clutch  54  connects the input shaft  17  with the sun gear member  22 , and the C 6  clutch  59  connects the common carrier member  29  with the transmission housing  60 . The sun gear member  22  rotates at the same speed as the input shaft  17 . The common carrier member  29  does not rotate. The ring gear member  24  rotates at the same speed as the ring gear member  34 . The ring gear member  24  rotates at a speed determined from the speed of the sun gear member  22  and the ring gear/sun gear tooth ratio of the planetary gear set  20 . The sun gear member  32  rotates at the same speed as the sun gear member  42 . The sun gear member  32  rotates at a speed determined from the speed of the ring gear member  34  and the ring gear/sun gear tooth ratio of the planetary gear set  30 . If the alternative single ring gear member  24 ′ is used in lieu of ring gear members  24  and  34 , then the speed of the single ring gear member  24 ′ is determined utilizing the speed of the sun gear member  22  and the ring gear/sun gear tooth ratio of the planetary gear set  20 . Additionally, the sun gear member  32  would then rotate at a speed determined from the speed of the single ring gear member  24 ′ and the ring gear/sun gear tooth ratio of the planetary gear set  30 . The ring gear member  44  rotates at the same speed as the output shaft  19 . The ring gear member  44  rotates at a speed determined from the speed of the sun gear member  42  and the ring gear/sun gear tooth ratio of the planetary gear set  40 . The numerical value of the Reverse Low speed ratio is determined utilizing the ring gear/sun gear tooth ratios of the planetary gear sets  20 ,  30  and  40 . 
     The first forward speed ratio is established with the engagement of the C 1  clutch  50  and the C 6  clutch  59 . The C 1  clutch  50  connects the input shaft  17  with the sun gear member  32 , and the C 6  clutch  59  connects the common carrier member  29  with the transmission housing  60 . The sun gear member  32  and the sun gear member  42  rotate at the same speed as the input shaft  17 . The common carrier member  29  does not rotate. The ring gear member  44  rotates at the same speed as the output shaft  19 . The ring gear member  44 , and therefore the output shaft  19 , rotates at a speed determined from the speed of the sun gear member  42  and the ring gear/sun gear tooth ratio of the planetary gear set  40 . The numerical value of the first forward speed ratio is determined utilizing the ring gear/sun gear tooth ratio of the planetary gear set  40 . 
     The second forward speed ratio is established with the engagement of the C 1  clutch  50  and the C 5  clutch  58 . The C 1  clutch  50  connects the input shaft  17  with the sun gear member  32 , and the C 5  clutch  58  connects the ring gear member  24  (or the single ring gear member  24 ′ in the event that the alternative embodiment is used) with the transmission housing  60 . The sun gear member  32  and the sun gear member  42  rotate at the same speed as the input shaft  17 . The ring gear member  24  and the ring gear member  34  (or in the case of the alternative embodiment, the single ring gear member  24 ′) do not rotate. The common carrier member  29  rotates at a speed determined from the speed of the sun gear member  32  and the ring gear/sun gear tooth ratio of the planetary gear set  30 . The ring gear member  44  rotates at the same speed as the output shaft  19 . The ring gear member  44 , and therefore the output shaft  19 , rotates at a speed determined from the speed of the common carrier member  29 , the speed of the sun gear member  42  and the ring gear/sun gear tooth ratio of the planetary gear set  40 . The numerical value of the second forward speed ratio is determined the ring gear/sun gear tooth ratios of the planetary gear sets  30  and  40 . 
     The third forward speed ratio is established with the engagement of the C 1  clutch  50  and the C 4  clutch  56 . The C 1  clutch  50  connects the input shaft  17  with the sun gear member  32 , and the C 4  clutch  56  connects the sun gear member  22  with the transmission housing  60 . The sun gear member  32  and the sun gear member  42  rotate at the same speed as the input shaft  17 . The sun gear member  22  does not rotate. The ring gear member  24  rotates at the same speed as the ring gear member  34 . The ring gear member  24  (or the single ring gear member  24 ′ in case the alternative embodiment is utilized) rotates at a speed determined from the speed of the common carrier member  29  and the ring gear/sun gear tooth ratio of the planetary gear set  20 . The ring gear member  34  (or, in the event that the alternative embodiment is utilized, the single ring gear member  24 ′), rotates at a speed determined from the speed of common carrier member  29 , the speed of the sun gear member  32  and the ring gear/sun gear tooth ratio of the planetary gear set  30 . The ring gear member  44  rotates at the same speed as the output shaft  19 . The ring gear member  44 , and therefore the output shaft  19 , rotates at a speed determined from the speed of the common carrier member  29 , the speed of the sun gear member  42  and the ring gear/sun gear tooth ratio of the planetary gear set  40 . The numerical value of the third forward speed ratio is determined utilizing the ring gear/sun gear tooth ratios of the planetary gear sets  20 ,  30  and  40 . 
     The fourth forward speed ratio is established with the engagement of the C 1  clutch  50  and the C 2  clutch  52 . The C 1  clutch  50  connects the sun gear member  32  with the input shaft  17 , and the C 2  clutch  52  connects the common carrier member  29  with the input shaft  17 . In this arrangement, all the members of the gear sets  20 ,  30  and  40  rotate at the same speed as the input shaft  17 . Thus the output shaft  19  rotates at the same speed as the input shaft  17  in a direct drive relationship. 
     The fifth forward speed ratio is established with the engagement of the C 2  clutch  52  and the C 4  clutch  56 . The C 2  clutch  52  connects the input shaft  17  with the common carrier member  29 , and the C 4  clutch  56  connects the sun gear member  22  with the transmission housing  60 . The common carrier member  29  rotates at the same speed as the input shaft  17 . The ring gear member  24  rotates at the same speed as the ring gear member  34 . The sun gear member  22  does not rotate. The ring gear member  24  (or the single ring gear member  24 ′ in the event that the alternative embodiment is utilized) rotates at a speed determined from the speed of the common carrier member  29  and the ring gear/sun gear tooth ratio of the planetary gear set  20 . The sun gear member  32  rotates at the same speed as the sun gear member  42 . The sun gear member  32  rotates at a speed determined from the speed of the ring gear member  34 , the speed of the common carrier member  29  and the ring gear/sun gear tooth ratio of the planetary gear set  30 . The ring gear member  44  rotates at the same speed as the output shaft  19 . The ring gear member  44 , and therefore the output shaft  19 , rotates at a speed determined from the speed of the common carrier member  29 , the speed of the sun gear member  42  and the ring gear/sun gear tooth ratio of the planetary gear set  40 . The numerical value of the fifth forward speed ratio is determined utilizing the ring gear/sun gear tooth ratios of the planetary gear sets  20 ,  30  and  40 . 
     The sixth forward speed ratio is established with the engagement of the C 2  clutch  52  and the C 5  clutch  58 . The C 2  clutch  52  connects the input shaft  17  with the common carrier member  29 , and the C 5  clutch  58  connects the ring gear member  24  with the transmission housing  60 . The common carrier member  29  rotates at the same speed as the input shaft  17 . The sun gear member  32  rotates at the same speed as the sun gear member  42 . The ring gear members  24  and  34  (or the alternative single ring gear member  24 ′) do not rotate. The sun gear member  32  rotates at a speed determined from the speed of the common carrier member  29  and the ring gear/sun gear tooth ratio of the planetary gear set  30 . The ring gear member  44  rotates at the same speed as the output shaft  19 . The ring gear member, and therefore the output shaft  19 , rotates at a speed determined from the speed of the common carrier member  29 , the speed of the sun gear member  42  and the ring gear/sun gear tooth ratio of the planetary gear set  40 . The numerical value of the sixth forward speed ratio is determined utilizing the ring gear/sun gear tooth ratios of the planetary gear sets  30  and  40 . 
     As set forth above, the engagement schedules for the torque-transmitting mechanisms is shown in the truth table of  FIG. 2 . This truth table also provides an example of speed ratios that are available utilizing the ring gear/sun gear tooth ratios given by way of example as follows N R1 /S R1 =3.86 and N R2 /S R2 =3.00 and N R3 /S R3 =3.86. N R1 /S R1  is the tooth ratio of planetary gear set  20 ; N R2 /S R2  is the tooth ratio of the planetary gear set  30 ; and N R3 /S R3  value is the tooth ratio of the planetary gear set  40 . It should be noted that single and double step ratio interchanges are of the single transmission variety. 
     The chart of  FIG. 3  describes the speed ratios and ratio steps that are obtained by the transmission of  FIG. 1  utilizing the same tooth ratios given above. For example, the step ratio between the first and second forward speed ratios is 1.714, while the step ratio between the Reverse Low and the first forward speed ratio is −1.29. A relatively wide ratio of 6.85 is obtained between the first and sixth forward speed ratios. 
     Shafting requirements for the transmission  14  are minimized by: (i) interconnecting the sun gear members  32  and  42  with the drum  92  and therefore the input shaft  17  via the inner shaft  95 , (ii) by selectively interconnecting the common carrier member  29  with the drum  92  and therefore the input shaft  17  via an intermediate shaft  94  when the C 2  clutch  52  is engaged, (iii) selectively interconnecting the sun gear member  22  with the input shaft  17  via the outer shaft  93  when the C 3  clutch  54  is engaged, and (iv) by creating the shafts  93 ,  94 ,  95  such that they are coaxially disposed. This allows for a compact arrangement. The C 3  clutch  54  is positioned radially inward of the C 1  clutch  50  and the C 2  clutch  52 . Because the C 3  clutch  54  rotates at higher speeds than the C 1  clutch  50  and the C 2  clutch  52 , spin losses are minimized by minimizing the radially displacement of the C 3  clutch  54  from a center axis of rotation (e.g., an axis defined by the input shaft  17  and the output shaft  19 ). 
     In the preferred embodiment, each of the pinion gears has a common number of teeth. For instance, in the transmission  14  of  FIG. 1 , the pinion gears  27 ,  28 ,  37 ,  47  and  48  may all have 27 teeth while sun gear member  22  has 21 teeth, sun gear member  32  has 27 teeth, sun gear member  42  has 21 teeth, ring gear member  24  and ring gear member  34  (or the alternative single ring gear member  24 ′) have 81 teeth and ring gear member  44  has 81 teeth. By providing pinion gears with a common number of teeth (i.e., having a predetermined size), bearings  21 A- 21 E may have a common bearing size and spindles  25 A- 25 E may be of a common size as well. 
     By utilizing a single common carrier member  29 , assembly time may be reduced as only one carrier member needs to be positioned within the transmission  14  rather than three separate members. The common carrier member  29  maybe located radially with precision on bushings  64 A and  64 B. The bushings  64 A and  64 B support inner radial portions of the common carrier assembly member  26  at a specifically designed radial and axial position. Accordingly, a method of assembling a transmission having multiple planetary gear sets includes providing a single carrier member  29  configured to rotatably support pinion gears  27 ,  28 ,  37 ,  47  and  48  for each of the planetary gear sets  20 ,  30  and  40 . The method further includes radially positioning the single carrier member within the transmission  14  on two axially spaced bushing  64 A,  64 B. 
     It is noted that the ratio coverage between the first and sixth forward speed ratios is 6.85 to 1 which provides a relatively high useable ratio coverage. The ratio of the forward and reverse speeds may be adjusted to provide nearly equal forward and reverse ratios. Other ratio coverages may be achieved with different gear tooth counts. For instance, low gear coverage may be reduced while the amount of overdrive may be increased. A low gear first forward speed ratio of 3.222 to a sixth forward speed ratio of 0.5 provides a total usable ratio coverage of 6.44 to 1. The final selection of ratio. coverage is based on cost, assembly and application guidelines. 
     The transmission  14  of  FIG. 1  provides pinion speeds, carrier speeds and clutch slip speeds compatible with very high engine input speeds, typical of smaller displacement, variable cam engines. With the selected tooth ratios discussed above, the C 3  clutch  54  will be characterized by speeds higher than those of the C 1  clutch  50  and the C 2  clutch  52 . Speed and torque calculations which will be readily understood by those skilled in the art (and which may be calculated based on gear tooth numbers) reveal that the transmission  14  provides very good torque sharing as the ratio steps progress, which improves durability of the transmission  14 . Additionally, those skilled in the art will readily understand that the transmission  14  is void of any internal power loops which enables a very high mechanical efficiency. 
     While the best modes for carrying out the invention have been described in detail, those familiar with the art to which this invention relates will recognize various alternative designs and embodiments for practicing the invention within the scope of the appended claims.

Technology Classification (CPC): 5