Interesting GM Patents

[Runge_Kutta]

3V-ENGINES (Twin and Single cam):
==============================

Twin cam:

<http://patft.uspto.gov/netacgi/nph-...&f=G&l=50&d=PALL&RefSrch=yes&Query=PN/6505592>

<http://patft.uspto.gov/netacgi/nph-...&f=G&l=50&d=PALL&RefSrch=yes&Query=PN/5560329>

Single Cam

<http://patft.uspto.gov/netacgi/nph-...&f=G&l=50&d=PALL&RefSrch=yes&Query=PN/6505589>

<http://www.uspto.gov/web/patents/patog/week02/OG/html/1266-2/US06505589-20030114.html>

Is this twin cam or single cam??

<http://www.corvetteenthusiast.com/toc200309.pdf>

TRANSMISSIONS:
===============

<http://patft.uspto.gov/netacgi/nph-...&f=G&l=50&d=PALL&RefSrch=yes&Query=PN/6468179>

<http://patft.uspto.gov/netacgi/nph-...&f=G&l=50&d=PALL&RefSrch=yes&Query=PN/6364804>

HYDROFORMING:
===============

<http://patft.uspto.gov/netacgi/nph-...&f=G&l=50&d=PALL&RefSrch=yes&Query=PN/6532639>

<http://www.uspto.gov/web/patents/patog/week11/OG/html/1268-3/US06532639-20030318.html>

<http://patft.uspto.gov/netacgi/nph-...&f=G&l=50&d=PALL&RefSrch=yes&Query=PN/5720092>

<http://patft.uspto.gov/netacgi/nph-...&f=G&l=50&d=PALL&RefSrch=yes&Query=PN/5557961>

<http://patft.uspto.gov/netacgi/nph-...&f=G&l=50&d=PALL&RefSrch=yes&Query=PN/5333775>

<http://nsmwww.eng.ohio-state.edu/THF_Automotive_Appln_Dr_G_Mississippi.pdf>

 

[Runge_Kutta]

Which do you like better; One or two cams??

<http://www.carcraft.com/techarticles/0303_GM/index3.html>
<http://www.uspto.gov/web/patents/patog/week02/OG/html/1266-2/US06505589-20030114.html>

 

[JBsC5]

Can you give your summary its somewhat involved

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Single cam three-valve engine overhead valve train


Abstract
A valve train layout for an overhead valve engine uses a single camshaft actuating two inlet valves and a single exhaust valve. The two inlet valves are preferably aligned on an inner side of a cylinder with the exhaust valve on an outer side. The inlet valves are preferably actuated by a cam follower and single push rod driving a rocker arm with dual arms and separate lash adjusters actuating the inlet valves. The exhaust valve is preferably actuated by a cam follower with a lash adjuster driving a single push rod operating a primary rocker arm that drives a secondary push rod pivoting an exhaust rocker arm to actuate the exhaust valve. In one embodiment, the primary rocker arm for the exhaust valve pivots on a common axis with the intake rocker arm. In an alternative embodiment, the primary rocker arm pivots on a canted pivot axis.


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Inventors: Hayman; Alan William (Romeo, MI); Mazzola, III; James Joseph (Dryden, MI); Rozario; Frederick J. (Fenton, MI); Black; Terry Wayne (Clarkston, MI)
Assignee: General Motors Corporation (Detroit, MI)
Appl. No.: 061684
Filed: February 1, 2002

Current U.S. Class: 123/90.23; 123/90.39; 123/90.61
Intern'l Class: F01L 001/26
Field of Search: 123/90.22,90.23,90.39-90.47,90.61-90.64



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References Cited [Referenced By]

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U.S. Patent Documents
4519364 May., 1985 Enomoto et al. 123/432.
4649874 Mar., 1987 Sonoda et al. 123/90.
5253620 Oct., 1993 Dohn et al. 123/90.
5398649 Mar., 1995 Shimamoto 123/90.
6205966 Mar., 2001 Breitenberger 123/90.

Primary Examiner: Lo; Weilun
Attorney, Agent or Firm: Hodges; Leslie C.

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Claims

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What is claimed is:

1. A valve train for a three-valve-per-cylinder engine, said valve train comprising:

a single camshaft extending longitudinally along a cylinder bank of the engine, the camshaft including at least one inlet cam and one exhaust cam for each engine cylinder;

at least one inlet cam follower engaging each of the inlet cams and an exhaust cam follower engaging each of the exhaust cams;

an inlet pushrod connecting each inlet cam follower with an inlet rocker arm;

a first exhaust pushrod connecting each exhaust cam follower with a first exhaust rocker arm;

a second exhaust pushrod connecting each first exhaust rocker arm with a second exhaust rocker arm; and

each inlet rocker arm operatively engaging at least one inlet valve for actuating the inlet valve and each second exhaust rocker arm operatively engaging at least one exhaust valve for actuating the exhaust valve.

2. A valve train as in claim 1 wherein each cylinder has dual inlet valves and a single exhaust valve.

3. A valve train as in claim 2 wherein said inlet valves are aligned longitudinally adjacent an inner side of the cylinder and the single exhaust valve is located adjacent an outer side of the cylinder.

4. A valve train as in claim 3 wherein said dual inlet valves are actuated by a single inlet rocker arm.

5. A valve train as in claim 4 wherein said second exhaust pushrod extends generally laterally across said cylinder between the first and second exhaust rocker arms to actuate said single exhaust valve.

6. A valve train as in claim 5 wherein said exhaust valve is offset longitudinally toward one side of the cylinder and an igniter is disposed longitudinally from the exhaust valve toward an opposite side of the cylinder.

7. A valve train as in claim 6 wherein a fuel injector is disposed adjacent the igniter in said opposite side of the cylinder.

8. A valve train as in claim 5 wherein said inlet rocker arm and said first exhaust rocker arm are pivotable on a common axis.

9. A valve train as in claim 5 wherein said inlet rocker arm and said first exhaust rocker arm are pivotable on non-parallel axes.
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Description

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TECHNICAL FIELD

This invention relates to valve trains for overhead valve engines and, in particular, to efficient valve trains for engines having three valves per cylinder.

BACKGROUND OF THE INVENTION

Internal combustion engines have been provided with numerous types of valve configurations, commonly including two, three or four valves. The valves are configured to obtain desirable gas flow, compression and combustion results while providing simplicity of the valve train to the extent possible for the particular engine arrangement. The most simple valve arrangements have generally involved a single intake valve and a single exhaust valve actuated by an overhead valve train including a single camshaft with cam followers and push rods actuating rocker arms of an overhead valve train. For obtaining higher engine outputs, valve configurations with four valves per cylinder have become common, utilizing dual intake valves on one side of the cylinder, usually the inside and dual exhaust valves on the other, outside, of the cylinder. Such valve configurations are more commonly actuated by multiple overhead camshafts directly driving the valves, or by a single overhead camshaft per cylinder bank driving some of the valves directly and others through finger followers or other rocker mechanisms.

An alternative valve configuration utilizes three valves per cylinder, generally including dual intake valves disposed along an inner side of the cylinder and a single exhaust valve located toward an outer side of the cylinder. Valves of this type may be actuated by overhead camshafts. Alternatively, a single camshaft per engine may connect through cam followers and push rods with rocker arms which actuate the various valves. The valve trains for such cylinder configurations vary in complexity and efficiency as do the valve arrangements themselves, and the relative placement of an igniter such as a spark plug and, if used, a fuel injector for the cylinder.

SUMMARY OF THE INVENTION

The present invention provides a valve train layout for an overhead valve (OHV) engine which uses a single camshaft in the cylinder block. A combustion chamber is provided for each cylinder which utilizes two inlet valves and a single exhaust valve. The two inlet valves may, as desired, be of equal or differing diameters. They are longitudinally aligned on the inner side of the cylinder (or combustion chamber) and are symmetrically located on opposite sides of a lateral plane through the cylinder axis and extending normal to the axis of the engine crankshaft and camshaft.

The single exhaust valve may be located either on the lateral plane or asymmetric to one side of the plane, depending on the requirements of the combustion system. Preferably, the exhaust valve is displaced to one side of the lateral plane and along the outer side of the cylinder and is inclined at a compound angle relative to the lower face of the cylinder head and the associated mounting deck of the engine cylinder block.

The inlet valves are preferably actuated by a cam follower and single push rod driving a rocker arm with dual arms carrying separate lash adjusters for actuating the two inlet valves. The exhaust valve is preferably actuated by a cam follower with a lash adjuster driving a single push rod operating a primary, or slave, rocker arm which in turn drives a secondary push rod actuating an exhaust rocker arm that opens the exhaust valve. Conventional springs are provided to close the valves and maintain the cam followers in contact with their cams.

In one embodiment, the primary or slave rocker arm for the exhaust valve is pivotable on a common axis with the intake rocker arm. In an alternative embodiment, the primary or slave rocker arm is pivotable on a canted pivot axis. In each case, a secondary push rod drives an exhaust rocker arm pivotable on a canted axis to engage the exhaust valve while pivoting in a plane common to the valve axis. In either embodiment, the camshaft may if desired be controlled by a cam phaser for varying valve timing.

Both embodiments for an overhead valve engine provide improved inlet port flow for better specific power output and dynamic range through the use of the two inlet valves. Improved exhaust port flow is obtained due to the compound angle of the exhaust valve. Better catalytic converter performance is obtained due to the reduced heat loss compared to a system with two or more exhaust valves. The compound angle of the exhaust valve also permits a larger exhaust valve for a given cylinder bore size. The space occupied by of the upper portion of the engine is reduced significantly over a multi-valve engine with overhead camshafts.

Hydraulic lash adjusters may be provided for all three valves in the engine layout although mechanical lash adjustment could be used if desired. With a camshaft phaser, the inlet and exhaust timing may be adjusted for better performance. A potential for reduced hydrocarbon emissions is possible due to reduced crevice volume associated with the single exhaust valve as opposed to dual exhaust valves in four-valve engines. The invention provides valve train driving inlet and exhaust valves on substantially different axes without using common rocker shafts that have longer, less desirable, rocker arm arrangements. Finally, a simple cam drive for a multi-valve engine cylinder concept is provided which allows improved cylinder head water jacket geometries and better cylinder head cooling performance.

These and other features and advantages of the invention will be more fully understood from the following description of certain specific embodiments of the invention taken together with the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a plan view of a single cylinder bank of an engine showing a first embodiment of valve train according to the invention;

FIG. 2 is a transverse cross-sectional view of a portion of the engine of FIG. 1 further illustrating the valve train embodiment;

FIG. 3 is a bottom view of the cylinder head of FIG. 1 showing the internal combustion chamber arrangement and valve locations;

FIG. 4 is a view similar to FIG. 2 showing an alternative embodiment of a valve train according to the invention; and

FIG. 5 is a view similar to FIG. 1 showing the alternative embodiment of FIG. 4.

DESCRIPTION OF THE PREFERRED EMBODIMENT

Referring first to FIGS. 1-3 of the drawing in detail, numeral generally indicates an internal combustion engine having a plurality of cylinders 12, only one of which is shown. The cylinders 12 are conventionally arranged in cylinder banks 14 with the upper ends of the cylinders being closed by cylinder heads 16.

The engine includes a cylinder block 18 supporting a crankshaft, not shown, which drives a camshaft 22 having a longitudinal axis extending parallel to the crankshaft axis, not shown. In a V-type engine, as shown, the camshaft 22 is located at the bottom of a V between the cylinder banks 14. FIGS. 1-3 illustrate the valve train arrangement in a single cylinder bank on one side of the engine and FIG. 2 additionally illustrates a portion of the valve train for a cylinder on the opposite cylinder bank 14 of the engine.

The valve train components include the camshaft 22 having a plurality of inlet cams 26. When rotated, each cam 26 reciprocates at least one inlet cam follower 28 connected with a push rod 30. The push rod actuates an inlet rocker arm 32 that rocks on a pivot axis 34 parallel to the camshaft axis 24. Rocker arm 32 includes a pair of actuating arms 36 each of which preferably carries a hydraulic lash adjuster 38. The lash adjusters engage a pair of inlet valves 40, each having a valve spring 42 for closing the respective valve and holding the cam follower 28 against the inlet cam 26. The inlet valves 40 are longitudinally aligned on the inner side of the cylinder and its associated combustion chamber 44 and are symmetrically located on opposite sides of a central lateral plane 46 extending normal to the crankshaft and camshaft axes and passing through the central axis, not shown, of the cylinder.

The camshaft 22 also carries a plurality of exhaust cams 48, only one of which is shown. Each exhaust cam actuates at least one cam follower 50 connecting with an exhaust push rod 52 for actuating a primary rocker arm or slave rocker arm 54 which in this embodiment is pivotable on the same pivot axis 34 as the inlet rocker arm 32. The cam follower and push rod for the rocker arm 54 illustrated are hidden in the drawing but are identical to the cam follower 50 and push rod 52 shown for actuating an exhaust valve in the opposite cylinder bank. The exhaust cam follower 50 preferably includes an internal hydraulic lash adjuster for adjusting the lash in the exhaust valve train driven by the follower.

The primary rocker arm 54 in turn engages a secondary push rod 56 which extends laterally at a slight longitudinally inward angle to engagement with a secondary rocker arm 58 pivotable on a canted axis 60. An actuating arm of rocker arm 58 directly engages the single exhaust valve 62 for its respective cylinder which includes a valve return spring 64 functioning as do the inlet valve springs to close the valve.

The exhaust valve may be located as desired on the outer side of the cylinder but as shown is desirably located asymmetrically longitudinally offset to one side of the cylinder, that is to one side of the central lateral plane 46. The exhaust valve 62 is angled at a compound angle with its axis directed generally toward the axis of the cylinder to provide a minimum crevice volume within the cylinder. Preferably, the rocking plane of the secondary rocker arm 58 where it contacts the exhaust valve 62 lies coplanar with the axis of the exhaust valve.

Spaced longitudinally on the other end of the outside of the cylinder from the exhaust valve location is an igniter in the form of a spark plug 66 or similar device. A fuel injector 68 may also be located at a position inwardly adjacent to the spark plug when the engine is provided with direct cylinder injection. However, the same cylinder configuration absent the direct fuel injector may also be utilized for an engine having manifold injection of the fuel.

In operation, rotation of the camshaft 24 during engine operation actuates the intake and exhaust valves in timed relation in accordance with predetermined valve lift diagrams. Actuation of the inlet cam follower 28 raises the push rod 30, pivoting the rocker arm 32 and opening both inlet valves 40 at the same time to provide a maximum amount of intake air flow. The separate lash adjusters at the ends of the actuating arms 36 of the rocker arm are required to separately adjust the lash of the two individual valves when driven by a single cam follower and push rod. If desired, the intake valves could be of differing sizes and could be driven by separate rocker arms actuated by individual push rods and cam followers from separate inlet cams in order to obtain a desired pattern of air flow within the cylinder. The general locations of the inlet valves being more or less symmetrical on the inner sides of the cylinder could remain the same.

Rotation of the camshaft 24 actuating the exhaust cam follower 50 with its integral lash adjuster raises the push rod 52 to actuate the primary rocker arm 54, secondary push rod 56 and secondary rocker arm 58 which opens the exhaust valve 62. Upon further rotation, the exhaust valve 62 is closed by the valve spring 64. Injection of fuel and ignition by the spark plug 66 will occur conventionally after compression of an air charge admitted through the intake valves with combustion and exhaust of the burned fuel-air charge following in conventional fashion.

FIGS. 4 and 5 illustrate an alternative embodiment similar to but modified from the embodiment of FIGS. 1-3 and wherein like numerals indicate like parts. Thus, engine 70 of FIGS. 4 and 5 includes a camshaft 22 and intake valve train like that of the first described embodiment. This alternative embodiment differs in that the slave rocker or primary exhaust rocker arm 72 is positioned on a canted axis 74 lying essentially parallel to the canted axis 60 of the secondary rocker arm 58. The two rocker arms are connected by a secondary push rod 76 essentially as before. As a result of repositioning the primary rocker arm 72, the exhaust primary push rod 52 and its connected follower 50 are also canted slightly from engagement of the follower with the exhaust cam 48 up to engagement of the push rod with the primary rocker arm 72. The configuration of the exhaust valve 62 and its valve spring 64 and the position of the spark plug 66 and the fuel injector 68, if used, are essentially the same as in the first described embodiment. Thus, the primary advantage of the second embodiment of FIGS. 4 and 5 is a more direct application of forces between the primary and secondary rocker arms due to the slight canting of the exhaust follower and push rod and the provision of a canted rocker arm pivot for the primary rocker arm.

While the invention has been described by reference to certain preferred embodiments, it should be understood that numerous changes could be made within the spirit and scope of the inventive concepts described. Accordingly, it is intended that the invention not be limited to the disclosed embodiments, but that it have the full scope permitted by the language of the following claims

 

[JBsC5]

Is this the new 6 speed automatic for C6?

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Six speed planetary transmission with two overdrive ratios


Abstract
A powertrain has a planetary transmission having a planetary gear arrangement and six torque transmitting mechanisms that are controlled to provide six forward ratios and a reverse ratio. The planetary gear arrangement includes two sun gear members, two ring gear members, and a planet carrier assembly member that are disposed in three axially spaced planes of gear mesh. The planet carrier member has a short pinion gear member, meshing (first plane) with one sun gear member, a first long pinion gear member having a small gear section, meshing (first plane) with the small pinion gear member and one ring gear member, and a large gear section, meshing(second plane) with the other sun gear member, and a second long pinion gear member, meshing (second and third planes) with both the large gear section (second plane) and the second ring gear member (third plane). The second long pinion gear member permits the pitch diameter of the second ring gear member to be radially inward of the pitch diameter of the first ring gear member.


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Inventors: Mathis; Roland (Ribeauville, FR); Andres; Marc (Schiltigheim, FR); Adolf; Pierre (Erstein, FR)
Assignee: General Motors Corporation (Detroit, MI)
Appl. No.: 853018
Filed: May 11, 2001

Current U.S. Class: 475/280; 475/338
Intern'l Class: F16H 003/62
Field of Search: 475/280,296,338



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References Cited [Referenced By]

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U.S. Patent Documents
2761333 Sep., 1956 Ravigneaux 475/280.
2800036 Jul., 1957 Miller 475/338.
5046999 Sep., 1991 Liu et al. 475/281.
5133697 Jul., 1992 Hattori 475/276.
5577976 Nov., 1996 Haka 475/280.
5911642 Jun., 1999 Andres et al. 475/10.
6364804 Apr., 2002 Haka 475/278.
Foreign Patent Documents
1042623 Nov., 1953 FR 475/280.
471394 Sep., 1937 GB 475/280.

Primary Examiner: Estremsky; Sherry
Attorney, Agent or Firm: Hodges; Leslie C.

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Claims

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What is claimed is:

1. A multi-speed planetary transmission comprising:

an input shaft for delivering power to the transmission;

an output shaft for delivering power from the transmission;

a planetary gear arrangement comprising a first sun gear member, a second sun gear member, a first ring gear member, a second ring gear member, and a planet carrier assembly member having a planet carrier rotatably supporting a plurality of pinion gear members including a first pinion gear member meshing with said first sun gear member, a second pinion gear member having a first gear section meshing with both said first pinion gear member and said first ring gear member and a second gear section meshing with said second sun gear member, and a third pinion gear member meshing with said second gear section and said second ring gear member, said first ring gear member being continuously connected with said output shaft;

a first torque transmitting mechanism selectively operatively connected between said input shaft and said first sun gear member;

a second torque transmitting mechanism selectively operatively connected between said planet carrier assembly member and said input shaft;

a third torque transmitting mechanism selectively operatively connected between said second sun gear member and said input shaft;

a fourth torque transmitting mechanism selectively operatively connected between said planet carrier assembly member and a transmission housing;

a fifth torque transmitting mechanism selectively operatively connected between said second ring gear member and said transmission housing;

a sixth torque transmitting mechanism selectively operatively connected between said second sun gear member and said transmission housing; and

said six torque transmitting mechanisms being selectively engaged in combinations of two to establish six forward ratios and one reverse ratio between said input shaft and said output shaft.

2. The multi-speed planetary transmission defined in claim 1 further comprising:

said second sun gear member and said second gear section being disposed in a gear plane axially intermediate gear planes containing said first ring gear member and said second ring gear member respectively.

3. A multi-speed planetary transmission comprising:

an input shaft for delivering power to the transmission;

an output shaft for delivering power from the transmission;

a planetary gear arrangement comprising a first sun gear member disposed in a first gear mesh plane, a second sun gear member disposed in a second gear mesh plane, a first ring gear member disposed in a first gear mesh plane, a second ring gear member disposed in a third gear mesh plane, and a planet carrier assembly member having a planet carrier rotatably supporting a plurality of pinion gear members including a first pinion gear member meshing with said first sun gear member in said first gear mesh plane, a second pinion gear member having a first gear section meshing with both said first pinion gear member and said first ring gear member in said first gear mesh plane and a second gear section meshing with said second sun gear member in said second gear mesh plane, and a third pinion gear member meshing with said second gear section in said second gear mesh plane and with said second ring gear member in said third gear mesh plane, said first ring gear member being continuously connected with said output shaft, and said second gear mesh plane being disposed axially between said first and third gear mesh planes;

a first torque transmitting mechanism selectively operatively connected between said input shaft and said first sun gear member;

a second torque transmitting mechanism selectively operatively connected between said planet carrier assembly member and said input shaft;

a third torque transmitting mechanism selectively operatively connected between said second sun gear member and said input shaft;

a fourth torque transmitting mechanism selectively operatively connected between said planet carrier assembly member and a transmission housing;

a fifth torque transmitting mechanism selectively operatively connected between said second ring gear member and said transmission housing;

a sixth torque transmitting mechanism selectively operatively connected between said second sun gear member and said transmission housing; and

said six torque transmitting mechanisms being selectively engaged in combinations of two to establish six forward ratios and one reverse ratio between said input shaft and said output shaft.

4. The multi-speed planetary transmission defined in claim 3 further comprising:

said first ring gear member having a pitch diameter and said second ring gear member having a pitch diameter numerically smaller than said pitch diameter of said first ring gear member.

5. The multi-speed planetary transmission defined in claim 3 further comprising:

said first gear mesh plane being disposed axially adjacent said output shaft.
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Description

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TECHNICAL FIELD

The present invention relates to power transmissions and more particularly to multi-speed power transmissions having six forward speed ratios.

BACKGROUND OF THE INVENTION

Multi-speed power transmissions are incorporated into powertrains to provide an extended operating range for the internal combustion engine that supplies the power for the powertrain. The number of ratios that are incorporated into automatic transmissions, especially for passenger car and small truck usage, has increased from two forward speed ratios to five forward speed ratios. Currently some manufacturers are considering transmissions incorporating six forward speed ratios. Examples of such proposals can be seen in the United States Patents issued to Liu and Malloy (U.S. Pat. No. 5,046,999) on Sep. 10, 1991 and Haka (U.S. Pat. No. 5,577,976) on Nov. 26, 1996.

These transmissions generally employ a Ravigneaux gear arrangement comprised of a long pinion and two or more short pinions that are intermeshing. The pinions interconnect two sun gears and two ring gears. By combining at least six selectively operable torque transmitting mechanisms (clutches and brakes), six forward speed ratios and one reverse ratio is available. The gears of these planetary arrangements are aligned in two axial planes. However the overdrive ratios achieved with these arrangements have a large step and the sixth ratio has a low numerical value.

Other six speed transmissions, having three axial planes of gears, have been proposed. One such mechanism is shown in U.S. Pat. No. 4,070,927 issued to Polak on Jan. 1, 1978. This arrangement uses three simple interconnected planetary gear sets and five torque transmitting mechanisms. The transmission described in the Polak patent has been used extensively in large trucks.

SUMMARY OF THE INVENTION

It is an object of the present invention to provide an improved multi-speed power transmission having six forward ratios and one reverse ratio.

In one aspect of the present invention, the transmission has a compound gear arrangement including two sun gear members and two ring gear members that are interconnected by a plurality of intermeshing pinion gears laid out in three axial planes. In another aspect of the present invention, at least one of the intermeshing pinion gears is a stepped pinion gear. In yet another aspect of the present invention, one of the ring gear members is an output member and the other ring gear member is a reaction member during two of the forward ratios.

In still another aspect of the present invention, the reaction ring gear member is smaller in diameter than the output ring gear member. In a further aspect of the present invention, the reaction ring gear member is disposed in one axial plane, the output ring gear member is disposed in another axial plane. In yet a further aspect of the present invention the planes containing the ring gear members are separated by an axial plane containing one of the sun gear members. In still a further aspect of the present invention, six torque transmitting mechanisms are employed to selectively provide six forward ratios, including two overdrive ratios, and a reverse ratio wherein the highest overdrive ratio is equal to or greater than 0.50.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic representation of a gearing arrangement incorporating the present invention.

FIG. 2 is a schematic representation of an end view of the gearing shown in FIG. 1.

FIG. 3 is a truth table of the engagement sequence of the torque transmitting mechanisms for the gearing arrangement.

FIG. 4 is an elevational view of the gearing arrangement incorporating the present invention and a portion of a transmission in which the gearing is assembled.

DESCRIPTION OF EXEMPLARY EMBODIMENTS

Referring to the drawings wherein like characters represent the same or corresponding components throughout the several views, there is seen in FIGS. 1 and 4 a powertrain 10 incorporating a conventional engine or prime mover 12, a conventional torque converter 14, a multi-speed power transmission 16 and a conventional final drive gearing 18. The final drive gearing 18 is drivingly connected to the drive wheels of a vehicle, not shown. The engine 12 and transmission 16 preferably include an electronic control module ECU, not shown, which ECU supplies control signals to the engine 12 and transmission 16 in response to operating parameter signals such as throttle position, engine speed, transmission input and output speeds, vehicle speed, and temperatures to name a few. The ECU typically includes a programmable digital computer that receives the operating signals and issues command to the engine 12 and transmission 16 to control fuel feed, fluid pressures, and transmission ratio interchanges.

The multi-speed transmission 16 includes a planetary gear arrangement 20, three rotating torque transmitting mechanisms or clutches 22, 24 and 26, and three stationary torque transmitting mechanisms or brakes 28, 30 and 32. The torque transmitting mechanisms 22, 24, 26, 28, 30, and 32 are conventional fluid operated selectively operable mechanisms which, when engaged, will transmit torque to and from the planetary gearing. The stationary torque transmitting mechanisms 28, 30, and 32 will transmit torque, when engaged, from the planetary gearing 20 to a transmission housing 34 while the rotating torque transmitting mechanisms 22, 24, and 26 will transmit torque, when engaged, from an input shaft 36 to the planetary gearing.

The planetary gearing arrangement 20 includes two sun gear members 38 and 40, two ring gear members 42 and 44, and a planetary carrier assembly member 46. The planetary carrier assembly member 46 has a planet carrier 48 on which is rotatably mounted a plurality of pinion gear members 50, 52 and 54. The pinion gear member 50 is a long pinion with two gear sections 50A and SOB that are equal diameter. In the alternative, the pinion gear member 50 can have the teeth formed continuously if this is more economical during manufacturing. The pinion gear member 52 is a long pinion with two gear sections 52A and 52B. Gear section 52A is of larger diameter than gear section 52B. The pinion gear member 54 is a short pinion that meshes between the sun gear member 40 and the gear section 52B. The gear section 52B also meshes with the ring gear member 44. The gear section 52A meshes with the sun gear member 38 and the gear section 50B. The gear section 50A meshes with the ring gear member 42. The gear meshes are distributed into three axial gear mesh planes P1, P2, and P3. The sun gear member 40, pinion gear member 54 and small gear section 52B are meshing in the plane P1. The sun gear member 38, large gear section 52A and the pinion gear member 50 are meshing in the plane P2. The ring gear member 42 and the pinion gear member 50 are meshing in the plane P3.

The torque transmitting mechanism 22 is selectively connectable between the input shaft 36 and the sun gear member 40. The torque transmitting mechanism 24 is selectively connectable between the input shaft 36 and the planetary carrier assembly member 46. The torque transmitting mechanism 26 is connectable between the input shaft 36 and the sun gear member 38. The torque transmitting mechanism 28 is selectively connectable between the planetary carrier assembly member 46 and the housing 34. The torque transmitting mechanism 30 is selectively connectable between the ring gear member 42 and the housing 34. The torque transmitting mechanism 32 is selectively connectable between the sun gear member 38 and the housing 34. The torque transmitting mechanisms 22, 24, 26, 28, 30, and 32 are selectively engageable in combinations of two to establish six forward speed ratios and one reverse ratio between the input shaft 36 and an output shaft 56 that is connected between the ring gear member 44 and the final drive gearing 18, as shown in FIG. 1.

As best seen in FIG. 4, the sun gear member 40 is splined to a shaft 58 that is operatively connected with the torque transmitting mechanism 22, not shown in FIG. 4. The planet carrier 48 of the planetary carrier assembly member 46 is drivingly connected with a sleeve shaft 60 that is operatively connected with the torque transmitting mechanism 24. The sun gear member 38 is splined to a sleeve shaft 62 that is operatively connected with a torque transmitting mechanism 32A and both the torque transmitting mechanism 32 and the torque transmitting mechanism 26, not shown. The torque transmitting mechanism 32A is disposed in parallel operating relation with the torque transmitting mechanism 32 and includes a plurality of friction plates 64 that are splined to an intermediate housing 66 which is secured in the transmission housing 34. The torque transmitting mechanism 32A also has a plurality of friction plates 68 that are splined to the outer race 70 of a one-way torque transmitting mechanism 72. An inner race 74 of the one-way mechanism 72 is drivingly connected with the sleeve shaft 62. A fluid operated piston 76 is slidably disposed in a housing 78 that is secured to the transmission housing 34 and also to the outer race 80 of a one-way torque transmitting mechanism 82. When the torque transmitting mechanism 32A is engaged, the sun gear member 38 is restrained from reverse rotation but is free to rotate in the forward direction. When the torque transmitting mechanism 32 is engaged, the sun gear member 38 is restrained from rotation in either direction.

The one-way mechanism 82 has an inner race 84 that is drivingly connected, through a sleeve shaft 86, with the planet carrier 48 of the planetary carrier assembly member 46 to restrain reverse rotation thereof. The sleeve shaft 86 is also connected with the torque transmitting mechanism 28 which includes a housing 88 that is secured to the housing 34 and in which is slidably disposed a piston 90. A plurality of friction plates 92 are drivingly connected with the housing 88 and a second plurality of friction plates 94 are drivingly connected with a hub 96 that is connected with the sleeve shaft 86. Whenever the torque transmitting mechanism 28 is engaged, the planet carrier 48 is restrained from rotation in both directions.

The housing 88 rotatably supports a hub 98 that is secured to the ring gear member 42 and is drivingly connected with a plurality of friction plates 100 that are components of the torque transmitting mechanism 30. The torque transmitting mechanism 30 also includes another plurality of friction plates 102, alternately spaced with the plates 100, and a piston member 104 that is slidably disposed in the housing 88. When the torque transmitting mechanism 30 is engaged by fluid pressure acting on the piston 104, the ring gear member 42 is restrained from rotation in both the forward and reverse directions. The ring gear member 44 is continuously connected with the output shaft 56 through a hub 106. As best seen in FIG. 4, the ring gear member 42 has a pitch diameter 108 that is disposed radially inward of a pitch diameter 110 of the ring gear member 44 which allows the present gearing arrangement to be installed in an envelope consistent in radial size with current four and five speed transmissions. The smaller pitch diameter of the ring gear member 42 and the stepped pinion gear member 52 also allows the overdrive ratios to have relatively large numerical ratios.

The one-way torque transmitting mechanisms 72 and 82 are employed to improve the shift feel between first and second and between second and third. These mechanisms are not necessary for the benefits derived from the present invention. With the use of electronic shift controls in the ECU, "clutch to clutch" shifts are achieved with a minimum of driveline disturbance. However, the use of the one-way torque transmitting mechanism, while adding mechanical content, will reduce the control function restraints imposed on the ECU. These features neither add to nor subtract from the present invention.

The truth table of FIG. 3 describes the engagement and interchange of the torque transmitting mechanisms for the forward and reverse ratios. To establish the reverse ratio, the torque transmitting mechanisms 26 and 28 are engaged. This establishes the sun gear member 38 as an input member and the planet carrier assembly member 46 as a reaction member. The numerical value of the reverse ratio is determined by the tooth ratio of the ring gear member 44, the sun gear member 38 and the ratio of the small gear section 52B to the large gear section 52A, hereinafter "pinion ratio".

The first forward ratio is established by the engagement of the torque transmitting mechanisms 22 and 28. Note that a forward/reverse interchange can be achieved by the swap of the torque transmitting mechanisms 22 and 26. In the first forward ratio, the sun gear member 40 is an input member and the planet carrier assembly member 46 is a reaction member. The numerical value of the first forward ratio is determined by the number of teeth on the sun gear member 40 and the ring gear member 44 (the gear meshes in plane 1). The first forward ratio is an underdrive ratio.

To establish the second forward ratio, the torque transmitting mechanism 28 is disengaged and the torque transmitting mechanism 30 is engaged. This interchange is controlled by the ECM. In the second forward ratio, the sun gear member 40 is the input member and the ring gear member 42 is a reaction member. The numerical value of the second forward ratio is determined by the tooth ratio of the ring gear member 42 to the sun gear member 40, the tooth ratio of the ring gear member 42 to the ring gear member 44, and the pinion ratio (the gear meshes in planes 1,2 and 3). The interchange from first to second is a single transition shift. The second forward ratio is an underdrive ratio.

To establish the third forward ratio, the torque transmitting mechanisms 30 and 32 are interchanged. The sun gear member 40 remains the input member and the sun gear member 38 is the reaction member. The numerical value of the third forward ratio is determined by the tooth ratio of the sun gear member 38 to the sun gear member 40, the sun gear member 28 to the ring gear member 44 and the pinion ratio (the gear meshes in planes 1 and 2). The interchange from second to third is a single transition shift. The third ratio is an underdrive ratio.

To establish the fourth forward ratio, the torque transmitting mechanisms 32 and 24 are interchanged. The planet carrier assembly member 46 and the sun gear member 40 are both input members and the planetary gear arrangement 20 is in a direct drive or one to one ratio. The interchange from third to fourth is a single transition shift.

To establish the fifth forward ratio, the torque transmitting mechanisms 22 and 32 are interchanged. This establishes the planet carrier assembly member 46 as an input member and the sun gear member 38 as a reaction member. The numerical value of the fifth forward ratio is determined by the tooth ratio of the sun gear member 38 to the ring gear member 44 and the pinion ratio (the gear meshes in planes 1 and 2). The fifth forward ratio is an overdrive ratio. The interchange from fourth to fifth is a single transition shift.

To establish the sixth forward ratio, the torque transmitting mechanisms 32 and 30 are interchanged in a single transition shift. The planet carrier assembly member 46 remains the input member and the ring gear member 42 becomes the reaction member. The numerical value of the sixth forward ratio is determined by the tooth ratio of the ring gear member 42 to the ring gear member 44 and the pinion ratio (the gear meshes in planes 1, 2, and 3). The sixth forward ratio is also an overdrive ratio.

The above interchanges are described for an upshift pattern. It will be apparent to those skilled in the art that the downshift pattern is the opposite and that all single step downshifts will be of the single transition type. The double step shifts (1.sup.st to 3.sup.rd, 2.sup.nd to 4.sup.th, 3.sup.rd 0 to 5.sup.th, and 4.sup.th to 6.sup.th) are all single transition also. It will also be appreciated that the pinion ratio affects all of the ratios except first forward and fourth forward.

The following are examples of the ratios that can be achieved with the present invention when the given gear tooth numbers are employed.
Gear No. of teeth No. of teeth
Member Example 1 Example 2
Sun gear member 38 36 33
Sun gear member 40 31 31
Pinion gear member 30 30
50 50A = 50B
Pinion gear member 27 28
54
Pinion gear member 42 45
52 gear section 52A
Pinion gear member 31 33
52 gear section 52B
Ring gear member 42 99 93
Ring gear member 44 109 112
Numerical Numerical
value Example value Example
Ratio established 1 2
1.sup.st forward 3.51 3.61
2.sup.nd forward 2.00 1.99
3.sup.rd forward 1.49 1.46
4.sup.th forward 1.00 1.00
5.sup.th forward 0.80 0.82
6.sup.th forward 0.59 0.62
Reverse 4.10 4.63

 

[JBsC5]

--------------------------------------------------------------------------------
Multi-speed power transmission having six forward ratios and one reverse ratio


Abstract
A powertrain has a planetary gear arrangement that is controllable to establish six forward speed ratios and one reverse ratio through the selective engagement of six torque transmitting mechanisms. The planetary gear arrangement includes three sun gear members, a ring gear member and a planet carrier assembly member having groups of three intermeshing pinion gear members axially arranged in three gear mesh planes. A first of the gear mesh planes has a sun gear member/pinion gear member mesh, a pinion gear member/pinion gear member mesh, and a ring gear member/pinion gear member mesh. A second of the gear mesh planes has a sun member/pinion gear member mesh, and at least one pinion gear member/pinion gear member mesh. A third of the gear mesh planes has a sun gear member/pinion gear member mesh. The ring gear member is continuously connected with a transmission output shaft. The torque transmitting mechanisms include three rotary devices that are selectively connectible between an input shaft and two of the sun gear members and the planet carrier assembly member, respectively; and three stationary devices that are selectively connectible between a transmission housing and the two of the sun gear members and the planet carrier assembly member, respectively. The torque transmitting mechanisms are engaged in combinations of two to establish three forward underdrive ratios, a direct forward ratio, two forward underdrive ratios, and a reverse ratio.


--------------------------------------------------------------------------------
Inventors: Haka; Raymond James (Brighton, MI)
Assignee: General Motors Corporation (Detroit, MI)
Appl. No.: 648761
Filed: August 28, 2000

Current U.S. Class: 475/278; 475/284; 475/288; 475/325
Intern'l Class: F16H 003/102; F16H 003/44
Field of Search: 475/278,279,280,281,284,288,313,325



--------------------------------------------------------------------------------

References Cited [Referenced By]

--------------------------------------------------------------------------------

U.S. Patent Documents
3733928 May., 1973 Uozumi et al. 74/753.
3859871 Jan., 1975 Uozumi et al. 74/761.
4070927 Jan., 1978 Polak 74/765.
5046999 Sep., 1991 Liu et al. 475/281.
5156575 Oct., 1992 Garrett 475/59.
5194056 Mar., 1993 Schiffhauer 475/325.
5495778 Mar., 1996 Mochizuki 74/606.
5577976 Nov., 1996 Haka 475/280.
5755637 May., 1998 Park 475/275.
6056659 May., 2000 McCarthy et al. 475/116.

Primary Examiner: Estremsky; Sherry
Assistant Examiner: Lewis; Tisha D.
Attorney, Agent or Firm: Hargitt; Laura C., Hodges; Leslie C.

--------------------------------------------------------------------------------

Claims

--------------------------------------------------------------------------------


What is claim is:

1. A powertrain having a multi-speed transmission comprising:

an input shaft for receiving power from a prime mover;

an output shaft for delivering power from said transmission;

a planetary gear arrangement having three sun gear members, a ring gear member and a planet carrier assembly member including at least three intermeshing pinion gear members rotatably mounted on a carrier member, each of said sun gears meshing with separate ones of said pinion gear members, said ring gear member meshing with one of said pinion gear members and being continuously connected with said output shaft;

a first torque transmitting mechanism selectively connectable between said input shaft and a first of said sun gear members;

a second torque transmitting mechanism selectively connectable between said input shaft and said planet carrier assembly member;

a third torque transmitting mechanism selectively connectable between said input shaft and a second of said sun gear members;

a fourth torque transmitting mechanism selectively connectable between a transmission housing and said planet carrier assembly member;

a fifth torque transmitting mechanism selectively connectable between said transmission housing and said second of said sun gear members;

a sixth torque transmitting mechanism selectively connectable between said transmission housing and a third of said sun gear members; and

said torque transmitting mechanisms being selectively engaged in combinations of two to establish three forward underdrive ratios, a direct drive ratio, two forward overdrive ratios, and a reverse ratio between said input shaft and said output shaft.

2. The transmission defined in claim 1 further comprising:

said planetary gear arrangement having three gear mesh planes;

one of said gear mesh planes including a sun gear member to pinion gear member mesh, a pinion gear member to pinion gear member mesh, and a pinion gear member to ring gear member mesh; and

another of said gear mesh planes having only a sun gear member to pinion gear member mesh.

3. A powertrain having a multi-speed transmission comprising:

an input shaft;

an output shaft;

a compound planetary gear arrangement having a first sun gear member, a second sun gear member, a third sun gear member, a ring gear member, and a planet carrier assembly member including a planet carrier having three pinion gear members rotatably mounted thereon with each pinion gear member meshing with one other pinion gear member and with respective ones of said sun gears, one of said pinion gear members also meshing with said ring gear member, and said ring gear member being continuously connected with said output shaft;

said sun gear members being axially spaced to define three gear mesh planes, a first of said gear mesh planes includes a meshing relation of said first sun gear member, a first and second of said pinion gear members, and said ring gear member, a second of said gear mesh planes includes a meshing relation of said second sun gear member, a third of said pinion gear members and at least said first of said pinion gear members, and a third of said gear mesh planes includes a meshing relation of said third sun gear member and said second of said pinion gear members; and

a plurality of torque transmitting mechanisms selectively operatively connected with respective ones of said sun gear members and said planet carrier assembly member and being engageable in combinations of two to establish a plurality of forward ratios and a reverse ratio between said input shaft and said output shaft.

4. The powertrain having a multi-speed transmission defined in claim 3 further comprising; said torque transmitting mechanisms including three rotary torque transmitting mechanisms and three stationary torque transmitting mechanisms;

said first sun gear member being operatively connected with a first of said rotary torque transmitting mechanisms;

said second sun gear member being operatively connected with a second of said rotating torque transmitting mechanisms and a first of said stationary torque transmitting mechanisms;

said planet carrier assembly member being operatively connected with a third of said rotary torque transmitting mechanisms and a second of said stationary torque transmitting mechanisms; and

said third sun gear member being operatively connected with a third of said stationary torque transmitting mechanisms.

5. The powertrain having a multi-speed transmission defined in claim 3 further comprising:

said second sun gear member having a smaller diameter than said first and third sun gear members; and

said third sun gear member having a larger diameter than said first sun gear member.

6. The powertrain having a multi-speed transmission defined in claim 3 further comprising:

said second sun gear member having a smaller diameter and less teeth than said first and third sun gear members; and

said third sun gear member having a larger diameter and more teeth than said first sun gear member.
--------------------------------------------------------------------------------

Description

--------------------------------------------------------------------------------


TECHNICAL FIELD

This invention relates to multi-speed power transmissions and more particularly to power transmissions having a compound planetary gear set controllable to provide six forward ratios and one reverse ratio.

BACKGROUND OF THE INVENTION

Multi-speed power transmissions are incorporated into powertrains to provide an extended operating range for the internal combustion engine that supplies the power for the powertrain. The number of ratios that are incorporated into automatic transmissions, especially for passenger car and small truck usage, has increased from two forward speed ratios to five forward speed ratios. Currently some manufacturers are considering transmissions incorporating six forward speed ratios. Examples of such proposals can be seen in the United States Patents issued to Liu and Malloy (U.S. Pat. No. 5,046,999) on Sep. 10, 1991 and Haka (U.S. Pat. No. 5,577,976) on Nov. 26, 1996.

These transmissions generally employ a Ravigneaux gear arrangement comprised of a long pinion and two or more short pinions that are intermeshing. The pinions interconnect two sun gears and two ring gears. By combining at least six selectively operable torque transmitting mechanisms (clutches and brakes), six forward speed ratios and one reverse ratio is available. The gears of these planetary arrangements are aligned in two axial planes. However the overdrive ratios achieved with these arrangements have a large step and the sixth ratio has a low numerical value.

Other six speed transmissions, having three axial planes of gears, have been proposed. One such mechanism is shown in U.S. Pat. No. 4,070,927 issued to Polak on Jan. 1, 1978. This arrangement uses three simple interconnected planetary gear sets and five torque transmitting mechanisms. The transmission described in the Polak patent has been used extensively in large trucks.

SUMMARY OF THE INVENTION

It is an object of the present invention to provide an improved multi-speed power transmission.

In one aspect of the present invention, a compound planetary gear set has three sun gear members, one ring gear member and a planet carrier assembly member. In another aspect of the present invention, the planet carrier assembly member has a pair of intermeshing long pinion gear members that interconnect two of the sun gear members and the ring gear member, and a short pinion gear member that interconnects the other sun gear member with one of the long pinion gear members. In yet another aspect of the present invention, the ring gear member is continuously connected with a transmission output shaft.

In still another aspect of the present invention, six selectively operable torque transmitting mechanisms are provided and engageable in combinations of two to establish six forward ratios and one reverse ratio. In yet still another aspect of the invention, the three sun gear members define the axial location of three gear planes with one of the gear planes containing one sun gear member, two long pinion members and the ring gear member. In a further aspect of the present invention, a second of the gear planes contains a second sun gear member, the short pinion gear member and the two long pinion gear members. In a yet further aspect of the present invention, a third of the gear planes contains the third sun gear member, and one of the long pinion gear members. In still a further aspect of the present invention, one of the long pinion gear members is engaged with other members in at least the first and third of the gear planes.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic representation of a powertrain including a compound gear arrangement incorporating the present invention.

FIG. 2 is a schematic representation depicting the meshing relationships of the gear members of the compound gear arrangement viewed in the axial direction.

FIG. 3 is a truth table describing the engagement and interchanges of the torque transmitting mechanisms that establish the ratios in the gear arrangement shown in FIG. 1.

DESCRIPTION OF AN EXEMPLARY EMBODIMENT

A powertrain 10, shown in FIG. 1, includes an engine 12, a torque converter 14, a multi-speed transmission 16, and a final drive 18. The engine 12, torque converter 14 and final drive 18 are conventional mechanisms well-known to those skilled in the art of power transmission in vehicles. The engine 12 and the transmission 16 preferably include a conventional electronic control unit (ECU), not shown, having a programmable digital computer. The ECU receives input signals from the engine 12, transmission 16 and the vehicle such as engine speed and torque, transmission input and output speed, temperature, and vehicle speed to name a few. The digital computer analyzes the signals and issues commands to control engine fuel feed, and transmission fluid pressure and ratio interchanges.

The transmission 16 has an input shaft 20, connected with the torque converter 14, an output shaft 22, connected with the final drive 18, a planetary gear arrangement 24, three rotary torque transmitting mechanisms 26, 28, and 30, and three stationary torque transmitting mechanisms 32, 34 and, 36. The final drive 18 is drivingly connected with the wheels, not shown, of the vehicle. The rotary torque transmitting mechanisms or clutches 26, 28, and 30 are each selectively connectable with the input shaft 20. The stationary torque transmitting mechanism or brakes 32, 34, and 36 are each selectively connectable with a transmission housing 38. The torque transmitting mechanisms 26, 28, 30, 32, 34, and 36 are preferably of the friction type with each having a fluid operated piston and a plurality of friction plates or discs. The stationary torque transmitting mechanism 32, 34, and 36 can, in the alternative, be constructed as band type devices having a fluid operated piston and a friction band surrounding a drum. There are other conventional torque transmitting mechanisms that can be employed also.

The planetary gear arrangement 24 includes three sun gear members 40, 42, and 44, a ring gear member 46, and a planet carrier assembly member 48. The planet carrier assembly member 48 has a planet carrier 50 that rotatably supports a short pinion gear member 52, and two long pinion gear members 54 and 56. The short pinion gear member 52 meshes with the long pinion gear member 54 which in turn meshes with the long pinion gear member 56. As is well-known, the meshing pinion gear members are arranged in three or more clusters in the planet carrier 50. The meshing pinion gear members have the radial orientation shown in FIG. 2.

The sun gear member 40 is selectively connectable with the input shaft 20 by the engagement of the torque transmitting mechanism 26. The planet carrier assembly member 48 is selectively connectable with the input shaft 20 by the engagement of the torque transmitting mechanism 28 and alternatively with the transmission housing 38 by the engagement of the torque transmitting mechanism 32. The sun gear member 42 is selectively connectable with the input shaft 20 by the engagement of the torque transmitting mechanism 30 and alternatively with the transmission housing 38 by the engagement of the torque transmitting mechanism 36. The sun gear member 44 is selectively connectable with the transmission housing 38 by the engagement of the torque transmitting mechanism 34. The ring gear member 46 is continuously connected with the output shaft 22 through a hub 58.

The sun gear member 40 meshes with the long pinion 54 to form a first gear mesh plane P1. The long pinion gear member 56 and the ring gear member 46 are also in mesh in the gear plane P1. The sun gear member 42, the short pinion gear member 52, and the long pinion gear member 54 are disposed in meshing relation in a second gear mesh plane P2. The sun gear member 44 and the long pinion gear member 56 are disposed in meshing relation in a third gear plane P3. The long pinion gear member 56 can have teeth formed continuously thereon such that it will mesh with the long pinion gear member 54 in the gear plane P2. However, manufacturing efficiency will probably dictate that the long pinion gear members 54 and 56 have teeth formed thereon only at the necessary points of mesh with other gear members. The long pinion gear member only needs to be in mesh at the gear planes P1 and P3.

To establish the reverse ratio in the planetary gear arrangement 24, the torque transmitting mechanisms 30 and 32 are engaged. The sun gear member 42 is an input member, the planet carrier assembly member 48 is a reaction member, and the ring gear member 46 is the output member. The numerical value of the reverse ratio is determined by the ratio of the number of teeth on the on the ring gear member 46 to the sun gear member 42. The gear meshes in the gear planes P2 and P1 are active.

To establish the first forward ratio, the torque transmitting mechanism 26 and 32 are engaged. A forward/reverse interchange can be achieved by the interchange of the torque transmitting mechanisms 30 and 26. During the first forward ratio, the sun gear member 40 is the input member, the planet carrier assembly member 48 is the reaction member, and the ring gear member 46 is the output member. The numerical value of the first forward ratio is determined by the ratio of the number of teeth on the ring gear member 46 to the sun gear member 40. The gear meshes in the gear plane P1 are active during this ratio which is an underdrive ratio.

To establish the second forward ratio, the torque transmitting mechanisms 32 and 34 are interchanged in a single transition shift. During the second forward ratio, the sun gear member 40 is the input member, the sun gear member 44 is the reaction member, and the ring gear member 46 is the output member. The numerical value of the second forward ratio is determined by the ratio of the number of teeth of the sun gear member 44 to the sun gear member 40, the sun gear member 44 to the ring gear member 46, and orbiting action of the planet carrier assembly member 48. The gear meshes in the gear planes P1 and P3 are active during this ratio which is an underdrive ratio.

To establish the third forward ratio, the torque transmitting mechanisms 34 and 36 are interchanged in a single transition shift. During the third forward ratio, the sun gear member 40 is the input member, the sun gear member 42 is the reaction member and the ring gear member 46 is the output member. The numerical value of the third forward ratio is determined by the tooth ratio of the sun gear member 42 to the sun gear member 40, the ring gear member 46 to the sun gear member 42, and the orbiting action of the planet carrier assembly member 48. The gear meshes in the gear planes P1 and P2 are active during this ratio which is an underdrive ratio.

To establish the fourth forward ratio, the torque transmitting mechanisms 36 and 28 are interchanged in a single transition shift. The engagement of both torque transmitting mechanisms 26 and 28 places the planetary gear arrangement 24 in a direct or one-to-one drive condition. The sun gear member 40 and the planet carrier assembly member 48 share the input member function and the ring gear member 46 is the output member.

To establish the fifth forward ratio, the torque transmitting mechanisms 26 and 36 are interchanged in a single transition shift. The planet carrier assembly member 48 is the input member, the sun gear member 42 is the reaction member, and the ring gear member 46 is the output member. The numerical value of the fifth forward ratio is determined by the ratio of the number of teeth on the sun gear member 42 to the ring gear member 46 and the orbiting action of the planet carrier assembly member 48. The gear meshes in the gear planes P1 and P2 are active during this ratio which is an overdrive ratio.

To establish the sixth forward ratio, the torque transmitting mechanisms 36 and 34 are interchanged in a single transition shift. The planet carrier assembly member 48 is the input member, the sun gear member 44 is the reaction member, and the ring gear member 46 is the output member. The numerical value of the sixth forward ratio is determined by the ratio of the number of teeth on the sun gear member 44 to the ring gear member 46 and the orbiting action of the planet carrier assembly member 48. The gear meshes in the gear planes P1 and P3 are active during this ratio.

The above described ratio interchanges define an upshift pattern. Those skilled in the art will recognize that the downshift pattern is the opposite. Also it will be apparent, from the above description and the truth table of FIG. 3, that all of the single step interchanges are of the single transition shift variety and that all of the double step (1.sup.st /3.sup.rd, 2.sup.nd /4.sup.th, 3.sup.rd /5.sup.th, and 4.sup.th /6.sup.th) interchanges are of the single transition shift variety. The numerical values of the transmission ratios are determined with the appropriate gear members having the values shown in the following table.
Gear member Sun gear 40 Sun gear 42 Sun gear 44 Ring gear 46
Number of 30 24 66 111
teeth

 

[Runge_Kutta]

My point here is to simply give people a glimpse of what is going on
at a technical level out of view of the customer. Patents aren't particularly
exciting reading - as you might have noticed. Interpretations are up
to the reader.

Engines
=======
<http://www.media.holden.com.au/division/2004_prodinfo/tech_displays/powertrain/OHV_3valve.pdf>
<http://www.carcraft.com/techarticles/0303_GM/index3.html>
<http://www.uspto.gov/web/patents/patog/week02/OG/html/1266-2/US06505589-20030114.html>
<http://www.corvetteenthusiast.com/toc200309.pdf>


RWD Transmission (X15R)
=======================
<http://www.media.holden.com.au/division/2004_prodinfo/tech_displays/powertrain/rwd_6speed.pdf>
<http://www.detnews.com/2003/autosinsider/0302/02/b02-73486.htm>
<http://www.sae.org/automag/techbriefs/12-2002/>
<http://www.newswire.ca/releases/December2002/18/c5039.html>
<http://www.newswire.ca/releases/December2002/18/c5045.html>

FWD Transmission (X22F)
=======================

<http://www.media.holden.com.au/divi...displays/powertrain/hydramatic_6speed_fwd.pdf>
<http://www.covisint.com/res/csm/articles/2003TransTech.shtml>

Hydroforming
============

<http://www.media.holden.com.au/divi...displays/tech_tools/aluminum_hydrofroming.pdf>
<http://www.autofieldguide.com/articles/050303.html>


"In 'A Design Concept for an Aluminum Sport Utility Vehicle Frame'
by Michael W. Danyo, Christopher S. Young, Henry J. Cornille, and Joseph Porcari
(SAE paper 2003-01-0572), the authors describe a study that was conducted under
the Partnership for a New Generation Vehicle (PNGV) program along with Alcan
Aluminum and The Budd Company: 'The specific objective of the study was
to assess the capability of an aluminum frame to achieve equivalent performance
to the 2002 Ford Explorer frame, but at a 40% weight reduction.' It wasn't
just a matter of determining whether they could devise a lighter frame, but
a frame that was actually compatible with other '02 Explorer elements,
like the body mounts, powertrain, closures, and so on.

The research concluded that it is possible to make a suitable frame, although
there would be both product and process changes required vis-à-vis switching
from steel to aluminum. For example, there would be an increase in the size
of the components, be it in thickness, section height or section width. Even
with the proposed size increases, the authors note, 'These dimensions will
yield a proposed design that yields a 44% weight savings for the frame relative
to the production steel frame while maintaining equivalent global torsion and
bending stiffness equal to or greater than the steel frame.' "

<http://www.sae.org/news/aluminum-suv.htm>

<http://www.sae.org/servlets/productDetail?PROD_TYP=PAPER&PROD_CD=1999-01-3180>

General Stuff
=============

http://www.media.holden.com.au/division/2004_prodinfo/tech_displays/

 

[Kotzenjunge]

(head explodes)

 

[Runge_Kutta]

Sorry if it is a bit much!!

What I think is really interesting is the fact that GM pursued two
distinct 3V configurations. I had no idea about the single cam version
until I started poking around the Patent Office website. Obviously,
Alan Hayman is a key driving Force here. By the way,
if you're willing to buy the patent documents, there are usually
diagrams and figures to look at rather than tedious legal-speak. As I look
at the diagrams for the single cam, 3V engine and compare it to the
picture on the GM brochure and the picture offered by Car Enthusiast
magazine, it all looks like the single cam engine and not the twin cam
engine. As to the transmissions, I expect that C6 will get the X15R
but there may be a few changes - especially the name. Maybe 6L60-E???
With regard to hydroforming, Bruggeman and his buddies have been making
these metal balloons for quite some time now and look to be rather good
at it.

Can 2V and 3V single cam engines use the same
short block??

 

[LongTimer]

Lots of great information to , but its giving me a :duh . My head's about to :blow .

 

[mrvette]

I dunno about the engines, but to ask.....to what specific advantage is this design that is worth the tooling costs??? or just a marketing play???

in the auto tranny department.....I feel the need for a 6 speed tranny is just silly allready....all that complexity and they still can't get a 4 speed tranny to last, what in hell can we expect out of a 6 speed??? forgetaboutit, untill they prove the point it's worth anything in durability (NOT) and weight savings (NOT) and fuel economy...highly doubtful, and finally performance.....then the 1/4 mile times are what would prove the point......looking at the ratios, seems to be a whole lot of shifting going on for what will be the slippage of the torque converter, with no real engine drop in RPMs....or is that the point??? making the engine rev to say 3 grand and stay there from stop to 80 mph??? more efficiency and hauling capacity from a smaller engine??? maybe......dunno.....interesting theory though.....but I have to say some years/decades ago, they had a thing called a dyno-flo....Buick I think, and that unit had many torque converters, or variable vane converters or some such arrangement of constantly varying the output speeds without clutch packs....
so just where is there any imporvement by using 6 speeds?? in the final anal-I -sis that is....

GENE

 

[JBsC5]

Six speed automatics increase fuel efficiency about 15% and can knock about .5-.7 tenths off 1/4 mile times...over 4 speed automatics..


In my opinion..its money well spent...although I prefer the DSG solution as it offers enthusiasts both automatic and six speed manual capability...(HEAVY TRAFFIC is still a ***** in manual after 2 or 3 hours at a shot)

 

[BlackNBlue95]

My head explode!

 

[wishuwerehere82]

It should be pointed out that the filing of a patent is not necessarily an indication of intent to manufacture, however there is a lot of investment in the invention and patenting process as well as the basic research involved.

The patent of the invention is considered intelectual property that can be used as a tool to bring suit against a competitor for using the idea in a competitive product.
It can also be bought and sold as a commodity if the inventing company decides not to pursue the invention for production.

As an inventor myself, I can tell you that a lot of time and effort is spent blocking competitors from using techniques that would otherwise cut into the profit margins of the inventing companies.
My .02