1. Field of the Invention
The present invention relates to an automatic transmission in the form of a belt-driven conical-pulley transmission, as known for example from DE 10 2004 015 215 and other publications, as well as a method for producing it and a motor vehicle equipped with it.
2. Description of the Related Art
Automatic transmissions in the broader sense are converters, whose momentary transmission ratio changes automatically, in steps or continuously, as a function of present or anticipated operating conditions, such as partial load and coasting, and environmental parameters, such as, for example, temperature, air pressure, and, humidity. They include converters that are based on an electrical, pneumatic, hydrodynamic, or hydrostatic principle, or on a principle which is a mixture of those principles.
The automation refers to a great variety of functions, such as start-up, choice of transmission ratio, or the type of transmission ratio change in various operating situations, where the type of transmission ratio change can mean, for example, shifting to different gear steps in sequence, skipping gear steps, and the speed of shifting.
The desire for convenience, safety, and reasonable construction expense determines the degree of automation, i.e., how many functions take place automatically.
As a rule, the driver can intervene manually in the automatic sequence, or can limit it for individual functions.
Automatic transmissions in the narrower sense, as they are used today primarily in the construction of motor vehicles, usually have the following structure:
On the input side of the transmission there is a start-up unit in the form of a regulatable clutch, for example a wet or dry friction clutch, a hydrodynamic clutch, or a hydrodynamic converter.
With a hydrodynamic converter or a hydraulic coupling, often a bridging clutch or lock-up clutch is connected parallel to the pump and turbine parts, which increases the efficiency by transferring the force directly and damps vibrations through defined slippage at critical rotational speeds.
The start-up unit drives a mechanical, continuously variable or stepped, multi-speed gearbox, which can include a forward/reverse driving unit, a main group, a range group, a split group, and/or a variable speed drive. Gearbox groups can be of intermediate gear or planetary design, with straight or helical tooth system, as a function of the requirements in terms of quietness of operation, space conditions, and transmitting options.
The output element of the mechanical transmission, a shaft or a gear, drives a differential directly or indirectly via intermediate shafts or an intermediate stage with constant transmission ratio, which can be configured as a separate gearbox or is an integral component of the automatic transmission. In principle, the transmission is suitable for longitudinal or transverse installation in the motor vehicle.
To adjust the transmission ratio in the mechanical transmission there are hydrostatic, pneumatic, and/or electrical actuators. A hydraulic pump, which operates on the displacement principle, supplies oil under pressure for the start-up unit, in particular the hydrodynamic unit, for the hydrostatic actuators of the mechanical transmission, and for lubricating and cooling the system. As a function of the necessary pressure and delivery volume, possibilities include gear pumps, screw pumps, vane pumps and piston pumps, the latter usually of radial design. In practice, gear pumps, vane pumps, and radial piston pumps have come to predominate for that purpose, with gear pumps and vane pumps offering advantages because they are less expensive to build, and the radial piston pump offering advantages because of its higher pressure level and better regulation ability.
The hydraulic pump can be located at any desired position in the transmission, on a main or a secondary shaft that is constantly driven by the drive unit.
Continuously variable automatic transmissions are known that consist of a start-up unit, a reversing planetary gearbox as the forward/reverse drive unit, a hydraulic pump, a variable speed drive, an intermediate shaft and a differential. The variable speed drive, in turn, consists of two pairs of conical disks and an endless torque-transmitting means. Each pair of conical disks includes a second conical disk that is movable in the axial direction. Between those pairs of conical disks passes the endless torque-transmitting means, for example a steel thrust belt, a tension chain, or a drive belt. Moving the second conical disk changes the running radius of the endless torque-transmitting means, and thus the transmission ratio of the continuously variable automatic transmission.
Continuously variable automatic transmissions (CVT) require a high level of pressure in order to be able to move the conical disks of the variable speed drive with the desired speed at all operating points, and also to transmit the torque with a sufficient base contact pressure with minimum wear.
In motor vehicles the need for comfort and convenience is generally very high, especially in regard to the noise level. The driver and passengers, especially in upscale vehicles, want there to be no disturbing noises coming from the operation of the vehicle's mechanical units. But the internal combustion engine, and also other mechanical units such as transmissions, does produce sounds, which can be widely perceived as disturbing. Thus, for example, in continuously variable transmissions where a plate-link chain is used there can be a sound, since such a plate-link chain, because of its construction with plate links and pins, produces a recurring impact due to the pins striking the conical disks of the transmission. In CVT transmissions, acoustic effects are generally attributed to the pin impact as the source. That acoustic excitation then produces resonances at the natural frequencies of the transmission housing (FE modes) or of the shafts (torsional modes, bending modes).
Another acoustic effect is produced by the CVT belt, the CVT band, or the CVT chain, which can vibrate on the tension side like a musical string; that can be suppressed for example by a slide bar. Torsional friction vibrations at frequencies of 10 Hz are known in clutches, for example, as grabbing. If the coefficient of friction gradient is such that the coefficient of friction decreases with increasing relative rotational speed or velocity, as the slippage changes, grabbing results. In automatic transmissions it is primarily the steel-to-paper coefficient of friction that is relevant.
Part of the purpose of the present invention is to improve the acoustics of such a transmission, and thus to improve the comfort—in particular the sound comfort —of a motor vehicle equipped with such a transmission. Another part of the purpose of the present invention is, after analyzing strong CVT vibrations and clarifying the associated operating mechanisms, to design appropriate countermeasures for minimizing—or if possible preventing—those vibrations, which lie for the most part in the acoustic range on the order of 400-600 Hz. Another part of the purpose of the present invention is to increase the endurance strength of components, and thus to prolong the operating life of such an automatic transmission. The reason for another part of the purpose of the present invention is to increase the torque transmission capability of such a transmission and to be able to transmit greater forces through the components of the transmission. Furthermore—hence that is another part of the purpose—it should be possible to economically produce such a transmission.
The parts of the problem are solved by the invention along with its refinements, presented in the claims and in the description, and are explained in connection with the drawing figures.
The analysis produces a simulation-based understanding of the nature of the vibration form, which involves a movement of the encircling chain coupled with a tipping or bending of the particular conical disk. The primary determinants of the frequency of the vibrations are the mass of the chain and the overall tipping and bending stiffness of the conical disks. That stiffness includes the inherent dishing of the disks, the tipping of the disks, the bending of the shafts as a result of their elasticity, and the tilt of the shafts as result of differences in bearing rigidities or bearing spacings. In addition, the coefficient of friction level and the gradient of the coefficient of friction, as well as the rotational speed and the transmission ratio, are determinants of the frequency.
Those findings are surprising, inasmuch as vibrations of the chain in the encircling arc, i.e., while it is being clamped in the disk set, have not been described before, and are also contrary to the view held heretofore that the frictional contact with the conical disks suppresses such vibrations in the arcs.
The influence of the CVT oil on such frictional vibrations has also not been described before, so that up until now those oils have been developed merely for friction that is high and is stable over time, as well as for low wear.
While it is known that with the movable CVT conical disks (movable disks) tilting play between the shaft and the movable disk has an effect on the efficiency, no vibrational bending, tilting, or wobbling motions of the movable disks have been described heretofore.
To solve that problem, it can therefore be necessary to consider more than one of the influenceable parameters, and thus, for example, to combine certain properties of the oil with certain mechanical configurations.
In accordance with the invention the problem is solved by a belt-driven conical-pulley transmission having pairs of conical disks on the input and output sides, each having a fixed disk and a movable disk, which are positioned in each case on shafts on the input side and on the output side, and are connectable by means of a endless torque-transmitting means for transmitting the torque, where at least one of the listed factors is optimized in terms of the acoustics of the transmission:
a viscous or hydraulic medium in the form of oil;
the surface quality of the contact regions between the conical disk and the endless torque-transmitting means;
the geometry of at least one conical disk;
the damping of at least one conical disk; and
the guidance of at least one conical disk.
It can be advantageous to use an oil having a coefficient of friction that is insensitive to the frictional speed. It can also be advantageous to optimize the contact surfaces between the conical disk and the endless torque-transmitting means, for example in regard to their topography.
Furthermore, it can be advantageous to provide at least one conical disk that is optimized for rigidity and/or at least one damped conical disk. It can also prove advantageous to integrate into the transmission at least one conical disk that is radially outwardly guided.
In addition, the present invention relates to a motor vehicle having a transmission in accordance with the invention.
The structure, operation, and advantages of the present invention will become further apparent upon consideration of the following description, taken in conjunction with the accompanying drawings in which:
In the illustration according to
Axially displaceable conical disk 5 can also be shifted to the left in the plane of the drawing in a known manner, where in that position plate-link chain 2 is in a radially inner position (which is given reference numeral 2a), producing a transmission ratio of belt-driven conical-pulley transmission 1 in the direction of a slower speed.
The torque provided by a drive engine, not shown in detail, is introduced into the input side part of the belt-driven conical-pulley transmission shown in
A torque introduced through gear 6 results in the formation of an angle of rotation between axially stationary spreader disk 11 and axially displaceable spreader disk 12, which results in an axial displacement of spreader disk 12 because of start-up ramps located on the latter, onto which the balls 14 run up, thus causing an axial offset of the spreader disks with respect to each other.
Torque sensor 10 has two pressure chambers 15, 16, of which first pressure chamber 15 is intended to be charged with a pressure medium as a function of the torque introduced, and second pressure chamber 16 is supplied with pressure medium as a function of the transmission ratio of the transmission.
To produce the clamping force that is applied as a normal force to plate-link chain 2 between axially stationary disk 4 and axially displaceable disk 5, a piston and cylinder unit 17 is provided which has two pressure chambers 18, 19. First pressure chamber 18 changes the pressure on plate-link chain 2 as a function of the transmission ratio, and second pressure chamber 19 serves in combination with torque-dependent pressure chamber 15 of torque sensor 10 to increase or reduce the clamping force that is applied to plate-link chain 2 between conical disks 4, 5.
To supply pressure medium, shaft 3 has three conduits 20, through which pressure medium is fed into the pressure chambers from a pump, which is not shown. The pressure medium is able to drain from shaft 3 through a drain conduit 21 on the outlet side, and can be conducted back to the circuit.
Applying pressure to pressure chambers 15, 16, 18, 19 results in a torque-dependent and ratio-dependent shifting of axially displaceable conical disk 5 on shaft 3. To seat shiftable conical disk 5, shaft 3 has centering surfaces 22, which serve as a sliding fit for displaceable conical disk 5.
As can be readily seen from
The reference numerals used in
In
In the area of the floor of that bore 24 the lateral bore 25 branches off; there can be a plurality of those arranged around the circumference. In the case shown, that lateral bore 25 is shown as a radial bore; however, it can also be produced at a different angle as an inclined bore. Bore 25 penetrates the outer surface of shaft 3 at a place which is independent of the operating state, i.e., for example independent of the transmission ratio setting, in an area which is always covered by movable disk 5.
By shifting lateral bore 25 to the zone covered by movable disk 5, shaft 3 can be made axially shorter, enabling construction space to be saved. In addition, shortening shaft 3 can also result in reduced strain.
The mouth of the conduit or lateral bore 25 can be located for example in the area of the groove 26, which is adjacent to the centering surface 22 of the shaft. That can be particularly advantageous if the tooth system 27, which connects movable disk 5 to shaft 3 so that it can be shifted axially but is rotationally fixed, is subjected to heavy loads, for example by the transmission of torque.
But in many cases the load on the tooth system 27 will not be the most critical design criterion, so that the mouth of bore 25 can be placed in the area of that tooth system, as shown in
In addition, in the
As explained earlier, with clutches, for example, a coefficient of friction that drops as the running or surface speed increases leads to grabbing, and hence to a decline in comfort. An effort should therefore be made to keep that decline in the coefficient of friction over the change of running or surface speed as small as possible.
The coefficient of friction gradient shown in
The spacing of the curves in the direction of the ordinate represents the scatter range of the coefficient of friction as a function of the clamping force or contact pressure. The bottom line represents a low contact pressure and the upper one in each case represents a higher contact pressure.
When comparing the former construction according to the upper graph and the embodiment according to the invention as shown in the lower graph, it is noticeable that at first the scatter range that is bounded by the two curves is smaller, resulting in a lesser dependence of the coefficient of friction on the contact pressure or clamping pressure existing at the time. Expressed in different terms, the embodiment according to the present invention (the lower graph) is less sensitive to changes in contact pressure.
It can also be seen from
Such a clearly defined pattern of the coefficient of friction over the range of running or surface speed and over the range of contact pressure, as shown in the lower graph of
The graphs in
The upper graph in
Investigations with simulations and measurements have shown that the vibration behavior, and hence the noise behavior, are influenced positively by an increased tilting stiffness of the axially movable disks, with that applying in particular, but not exclusively, in regard to the movable disk on the output side. In general it has turned out that an increased bending stiffness, whereby the opening of the conical disks when under load is reduced, especially of the set of conical disks on the output side, the vibration amplitude, which is significant in regard to the noise, is lessened. A comparable effect can be achieved through increased damping at that location.
The movable disk 33 shown in
Movable disk 33 according to
To stiffen movable disk 33 in the axial direction, a stiffening collar can also be applied radially at the outside, as shown in
In
f and 5g show a stiffening of the connection of the disk to the shaft. Here, first of all, hub 37 of movable disk 33 is connected to the radially outwardly extending part of movable disk 33 by means of a stiffening ring 38, so that a deformation of that area is at least reduced. Furthermore, there are again radial stiffening ribs 34, which are connected on one side to stiffening ring 38 and on the other side to hub 37 of movable disk 33.
a through 6e show the principles of damping possibilities for the axially moving disk or movable disk 33 on the output side, which are also applicable, however, to the axially moving disk or movable disk 5 on the input side.
a shows first of all a subdivision of hub 37 into individual lamellae. That bundle of lamellae is pressed together by the clamping pressure that is applied through the hydraulic medium and thus produces a damping effect.
In
d and 6e both show springs 39, which increase the friction between the individual cylinders of lamellae through additional radial clamping pressure, which simultaneously increases the damping effect. It would also be possible in
f and 6g show a different approach to a solution, which involves changing the direction of tilt of the movable disk. With the usual guidance of the movable disk by its radial inner region or by its hub 37, the radial outer region of that movable disk shows the greatest deflection in the direction of tilting. To counter that, it is possible in principle to guide the movable disk at the outside, so that its radially outer regions lie against the outer guide 40 and hence cannot deflect there. Tilting would then occur at the radially inner region of movable disk 33, against which countermeasures could again be taken as described above. In that case, care must be taken, however, to avoid jamming or clamping of movable disk 33 between the guides.
Although particular embodiments of the present invention have been illustrated and described, it will be apparent to those skilled in the art that various changes and modifications can be made without departing from the spirit of the present invention. It is therefore intended to encompass within the appended claims all such changes and modifications that fall within the scope of the present invention.
Number | Date | Country | Kind |
---|---|---|---|
10 2004 040 826.2 | Aug 2004 | DE | national |
10 2004 041 715.6 | Aug 2004 | DE | national |
10 2004 042 883.2 | Sep 2004 | DE | national |
10 2004 043 536.7 | Sep 2004 | DE | national |
10 2004 044 190.1 | Sep 2004 | DE | national |
10 2004 046 213.5 | Sep 2004 | DE | national |
This application claims the benefit of U.S. Provisional Application Ser. No. 60/662,424, filed on Mar. 16, 2005.
Number | Date | Country | |
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60662424 | Mar 2005 | US |