Abstract:
An automatic transmission in the form of a belt-driven conical-pulley transmission having conical disk sets on the power input and power output sides, and an endless torque-transmitting member interconnecting the input side and the output side disk sets for transmitting torque therebetween. At least one stop at at least one of the end positions of at least one axially displaceable disk is provided with a cushioned retarding component for minimizing disk impact loads.

Description:
BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     The present invention concerns an automatic transmission in the form of a belt-driven conical-pulley transmission, such as is known, for example, from DE 10 2004 015 215 and other publications, as well as to a method for controlling and/or operating it and a 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 thrust and environmental parameters, such as temperature, air pressure, and humidity. They include converters that are based on an electrical, pneumatic, hydrodynamic, or hydrostatic principles, or on a principle that is a mixture of those principles. 
     The automation refers to a variety of functions, such as start-up, choice of transmission ratio, or type of transmission ratio change in various operating situations, where the type of transmission ratio change can mean, for example shifting, to different gear levels in sequence, skipping gear levels, and the speed of shifting. 
     The desire for convenience, safety, and reasonable construction expense determines the degree of automation, i.e., how many functions run 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, often a lock-up clutch is connected in parallel with the pump and turbine parts, which increases the efficiency by transmitting the force directly and damps the oscillations through defined slippage at critical rotational speeds. 
     The start-up unit drives a mechanical, continuously adjustable or stepped multi-speed gearbox, which can include a forward/reverse drive unit, a main gear group, a range gear group, a split gear group, and/or a variable speed drive unit. Transmission gear groups can be of intermediate gear or planetary design, with spur or helical tooth system, as a function of the requirements in terms of quietness of operation, space conditions, and transmission ratio 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 a constant transmission ratio. The differential can be configured as a separate gearbox or as an integral component of the automatic transmission. In principle, the transmission is suitable for longitudinal or transverse installation in the vehicle. 
     To adjust the transmission ratio in the mechanical transmission there are provided hydrostatic, pneumatic, and/or electrical actuators. A hydraulic pump, which operates by 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. Depending upon 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 and radial piston pumps have come to predominate for that purpose, with gear pumps offering advantages because they are less expensive to build, and radial piston pumps offering advantages because of their higher pressure level and better ability to be regulated. 
     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 adjustable 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 unit, an intermediate shaft and a differential. The variable speed drive unit in turn is made up of two pairs of conical disks and an endless torque-transmitting means. Each pair of conical disks includes a second conical disk that is displaceable in the axial direction. Between the pairs of conical disks runs the endless torque-transmitting means, which can be, for example, a thrust element belt, a tension chain, or a strap belt. Moving the second conical disk changes the running radius of the endless torque-transmitting means, and thus the transmission ratio of the continuously adjustable automatic transmission. 
     Continuously adjustable automatic transmissions require a high pressure level in order to be able to adjust the conical disks of the variable speed drive unit with the desired speed at all operating points, and also to transmit the torque with sufficient basic pressure with minimum wear. 
     One object of the present invention is to increase the operational strength of the components and thus to prolong the operating life of such an automatic transmission. A further object of the present invention is to increase the torque transmission capability of such a transmission and to be able to transmit higher forces through the components of the transmission. Furthermore, hence another object, such a transmission should be cost-effectively producible. 
     Another object of the invention is to be able to move the variable speed drive unit of a belt-driven conical-pulley transmission at the highest possible speed, i.e., to achieve the maximum adjustment between underdrive and overdrive, or vice versa, as quickly as possible. In the previously known continuously variable vehicle transmissions, in particular in transmissions having an endless torque-transmitting means, the quick adjustment proceeds in most cases in such away that a quick transmission ratio adjustment is typically transmission ratio regulated. But that transmission-ratio-regulated operation must take account of the former type of regulation itself. In order to prevent oscillations, one must normally put up with sacrifices in the adjustment dynamics, i.e., a slowing of the adjustment, because for reasons of stability the maximum force cannot be utilized during the entire adjustment process. 
     Furthermore, the speed of adjustment must be reduced before the stops are reached, since the latter cannot be approached with high dynamics, and are also unable to assume the necessary retardation process themselves. In particular in transmissions such as, for example, a concept in which a CVT transmission is operated in combination with an automated shift transmission or a stepped automatic transmission that is connected in series with it, it is important when shifting to be able to adjust the CVT part of the transmission as quickly as possible. 
     The several objects are achieved by the invention presented in the claims and the description and explained in connection with the figures, along with its refinements. 
     SUMMARY OF THE INVENTION 
     In accordance with the invention a contribution is made to solving the problem and to improving known transmissions by a belt-driven conical-pulley transmission having pairs of conical disks on the power input side and the power output side. The disk pairs each have a fixed disk and a displaceable disk, which are positioned on respective shafts on the input side and on the power output side and are connectable by means of an endless torque-transmitting means. At least one stop is provided with a retarding mechanism at least at one of the end positions of at least one displaceable disk. 
     That approach leads to optimization of the stop or stops. It is advantageous if they are designed in such away that shortly before the limit stop, at which the endless torque-transmitting means, for example in the form of a chain, can run out of the disk set, elasticity is built in which ensures that the disk set sustains no damage by travel to the stop. 
     It can be especially advantageous in a belt-driven conical-pulley transmission in accordance with the invention, if the retarding mechanism operates automatically. 
     It can be advantageous in a belt-driven conical-pulley transmission in accordance with the invention if the stop has a damping ring, which can be designed in more than one piece. 
     In general, it can be advantageous if the damping ring is made of a steel material that has some elasticity. 
     In addition, the damping ring can be enclosed by two bearing shells. 
     In a belt-driven conical-pulley transmission in accordance with the invention, it can be advantageous if a pressure medium is compressed in the area of the stop for the hydraulic medium present in the belt-driven conical-pulley transmission to be used as the pressure medium. 
     To that end, the hydraulic medium needed for adjusting the transmission ratio can be used as the pressure medium. 
     It is possible, for example, to bring the hydraulic medium through a cover to a specially formed stop, and of forming that stop in such a way that during the adjustment and shortly before the stop compression of the hydraulic medium occurs and damping is thereby achieved. To that end, the oil needed for the adjustment can be fed, for example, through a separate branch pipe to the damping stop. 
     It can be especially advantageous in a belt-driven conical-pulley transmission in accordance with the invention if the maximum adjustment force is generated to adjust the transmission ratio, whereby it can be useful to assist the quick adjustment of the maximum adjusting force by means of software, which can be achieved, for example, by an offset in the control variable, wherein the offset can be an electrical current offset. 
     The software control can occur, for example, in that at the command “quick adjustment” the control variable is manipulated by an offset value, precontrolled, in such a way that the maximum adjusting force is generated for the adjustment. It is conceivable, for example, in an underdrive quick adjustment for the current to then be immediately increased by an offset to 1000 mA, for example. Shortly before the stop is reached, the current can then be lowered to a value that is necessary to hold the underdrive transmission ratio. In most cases that is known when the particular load situation (variable speed drive unit torque, speed of rotation) is known. 
     It would also be possible, if one wished to remain in a quasi-regulated mode, to provide for a regulator parameter shift or a regulator structure shift for the quick adjustment, so that a maximum adjusting force can be produced in the shortest possible time. At the same time, the set point should be changed so that a large control deviation is achieved, thereby changing the control variable so that maximum adjustment forces are achieved. 
     It can also be advantageous with a belt-driven conical-pulley transmission in accordance with the invention if during a quick adjustment the regulated operating condition is supported by having values produced directly for the set point and that result in a large control deviation; the control variable can assume extreme values directly. 
     It can also be advantageous if extreme values are applied for the control variable in a very short time. 
     In a belt-driven conical-pulley transmission in accordance with the invention it can be advantageous if the quick adjustment is supported by the software in such a way that an increase or changeover of the regulating parameters occurs in transmission-ratio-regulated operation, so that the control values assume high values or extreme values in a short time, whereby a structural change can take place in the regulator so that the control values assume high values or extreme values in a short time. 
     The invention also relates to a method for controlling and/or operating a transmission in accordance with the invention. 
     In addition, the invention relates to a vehicle having a transmission in accordance with the invention. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       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: 
         FIG. 1  is cross-sectional view of a part of a belt-driven conical-pulley transmission; 
         FIG. 2  is a cross-sectional view of a part of another embodiment, corresponding substantially to  FIG. 1 ; 
         FIG. 3  shows exemplary embodiments of output side pairs of conical disks; 
         FIG. 4  is an enlarged view of a portion of the hub area of the output side displaceable disk; 
         FIG. 5  is a partial section of an embodiment of a damping ring; and 
         FIG. 6  is another enlarged view of a portion of the hub area of an output side displaceable disk. 
     
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Referring now to the drawings,  FIG. 1  shows only part of a belt-driven conical-pulley transmission, namely the input or driven side of the belt-driven conical-pulley transmission  1 , which is driven by a drive engine, for example an internal combustion engine. In a fully implemented belt-driven conical-pulley transmission, assigned to that input-side part is a complementarily-designed output side part, the two parts being connected by an endless torque-transmitting means in the form of a plate-link chain  2 , for example, for transmitting torque. Belt-driven conical-pulley transmission  1  has a shaft  3  on its input side, which is designed in the illustrated exemplary embodiment as integrally formed with a stationary conical disk or fixed disk  4 . That axially fixed conical disk  4  is positioned in the axial longitudinal direction of shaft  3  close to and opposite an axially displaceable conical disk or displaceable disk  5 . 
     In the illustration in accordance with  FIG. 1 , plate-link chain  2  is shown in a radial outer position on disk pair  4 ,  5  on the input side, resulting from the fact that the axially displaceable conical disk  5  is repositioned toward the right in the drawing, and that repositioning movement of axially displaceable conical disk  5  results in a movement of plate-link chain  2  in the radially outward direction, producing a change in the transmission ratio of the transmission toward greater speed. 
     Axially displaceable conical disk  5  can also be repositioned 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 indicated by reference numeral  2   a ), 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) is introduced into the input side part of the belt-driven conical-pulley transmission shown in  FIG. 1  by way of a gear  6  mounted on shaft  3 . Gear  6  is supported on shaft  3  by means of a roller bearing in the form of a ball bearing  7  that absorbs axial and radial forces, and is fixed on shaft  3  by means of a washer  8  and a shaft nut  9 . Between gear  6  and axially displaceable conical disk  5  is a torque sensor  10 , with which a spreader disk configuration  13  having an axially fixed spreader disk  11  and an axially displaceable spreader disk  12  is associated. Located between the two spreader disks  11 ′  12  are roller elements, for example in the form of the illustrated balls  14 . 
     A torque introduced via 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 repositioning 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 provided 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 the sides of 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 disks  4 ,  5 . 
     To supply pressure medium to the pressure chambers, shaft  3  has three conduits  20 , through which pressure medium is fed from a pump (not shown). The pressure medium can drain from shaft  3  through a conduit  21  on the outlet side, and can be conducted back to the circuit. 
     The pressurization pressure chambers  15 ,  16 ,  18 ,  19  results in a torque-dependent and transmission-ratio-dependent repositioning of axially displaceable disk  5  on shaft  3 . To receive displaceable disk  5 , shaft  3  has centering surfaces  22 , which serve as a sliding seat for displaceable disk  5 . 
     As can be readily seen from  FIG. 1 , in the area of each of the bearing positions of disk  5  on shaft  3 , belt-driven conical-pulley transmission  1  has a noise damping device  23 . For that purpose the noise damping device can have a ring body and a damping insert, or can consist only of a damping insert. 
     The reference numerals used in  FIG. 1  also apply to the essentially comparable features in the other drawing figures. Thus the drawing figures are to be regarded as a unit in that respect. For the sake of clarity, only the reference numerals that go beyond those in  FIG. 1  are used in the other figures. 
     In  FIG. 2 , only the middle one of the three conduits  20  is fully configured in a form that is modified from that shown in  FIG. 1 . It is evident that bore  24  that forms the central conduit  20  is produced as a blind bore from the side shown on the right in  FIGS. 1 and 2 , and is significantly shorter in  FIG. 2  than in  FIG. 1 . Such blind bores are complex and expensive to produce and require a very high degree of precision in manufacturing. The expense of production and the requirements in terms of process reliability increase disproportionately with the length. Thus, shortening a bore of that sort has a favorable effect on the production costs, for example. 
     In the area of the end of bore  24  the lateral bore  25  branches off, of which there can be a plurality arranged circumferentially. In the case shown, 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 shell of shaft  3  at a place that is independent of the operating mode, i.e., independent of the transmission ratio setting, for example, in a region that is always covered by displaceable disk  5 . 
     By shifting lateral bore  25  to the region covered by displaceable 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 loading. 
     The outlet of the conduit or lateral bore  25  can be located in the region of the recess  26 , for example, which is adjacent to the centering surface  22  of the shaft. That can be particularly advantageous if the tooth system  27 , which connects displaceable disk  5  to shaft  3  so that it can be axially displaced but is rotationally fixed, is subjected to heavy demands, 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 outlet of bore  25  can be placed in the area of that tooth system, as shown in  FIG. 2 . Placing lateral bore  25  at the tooth system  27  region instead of in the recess  26  produces an advantage through the fact that a greater resistance moment is present, which reduces the bending stress in the surface layer. In addition, the surface moment of inertia is greater at that location, while the critical layer, which is disturbed by lateral bore  25 , remains at an approximately constant radius. That results in a significant reduction of the stresses in the critical area around the outlet of lateral bore  25  between the teeth of tooth system  27 . 
     The system for supplying hydraulic fluid is identical in  FIGS. 1 and 2 , since pressure chambers  15  and  19  are connected to each other and displaceable disk  5  has connecting bores  28 , which connect the region of the tooth system  27  with pressure chamber  19 . 
     In the figures, displaceable disk  5  is in its most extreme left position, which corresponds to the start-up transmission ratio or underdrive. If displaceable disk  5  is now repositioned to the right in the direction of fixed disk  4 , there is always part of the hollow or of chamber  29  over the outlet of the lateral bore or of conduit  25 , so that the necessary fluid supply is always ensured, just as in  FIG. 1 . Also as in  FIG. 1 , there are two shift states for pressure chamber  16 , which depend upon the axial position of displaceable disk  5 . In the illustrated position the control bores  30  are free, so that the conduit  20  that is connected to them and is closed axially with a stopper  31 , and the pressure chamber  16 , which is connected to the latter through a conduit (not shown), are not pressurized or have only ambient pressure. If displaceable disk  5  is now moved toward fixed disk  4 , it passes over control bores  30 , so that starting at a certain distance chamber  29  comes to rest over the openings of control bores  30 . In chamber  29 , however, a high pressure as a function of the torque prevails, which is then also conveyed through control bores  30  and conduit  20  into pressure chamber  16 , so that high pressure is also present there. In that way two shift states are realized, which control the clamping force as a function of the transmission ratio. 
     In addition, in  FIG. 2  there is a disk spring  32  that moves displaceable disk  5  to a predetermined axial position when transmission  1  is not under pressure, enabling a transmission ratio of transmission  1  to be set that prevents excessive loads, for example when the vehicle is being towed. 
       FIG. 3  shows two possible configurations of conical disk set  33  on the output side, with the lower half of the figure showing a disk set constructed in accordance with the single piston principle, while the upper half shows a disk set constructed in accordance with the dual piston principle, as described, for example, in DE 103 54 720.7. 
     In the dual piston version, separate pistons are available for the clamping and the repositioning, whereas in the single piston version only one piston/cylinder unit introduces the corresponding force into the disk set. 
     Compared to the customary versions heretofore, spring  34  here has a larger diameter, so that its point of application on displaceable disk  35  is radially farther outward. One of the advantages resulting from that arrangement is that more construction space is available to thicken the conical disk neck or hub  36 , or to design it stronger geometrically and increase its diameter. The resulting gain in strength was already described earlier. In the dual piston version shown in the upper half of  FIG. 3 , it results in a modified arrangement of spring  34  to the effect that it is repositioned from the radially inner pressure chamber into the radially outer pressure chamber. 
     Sheet metal part  37  that radially inwardly supports spring  34  is firmly connected to displaceable disk  35 , and its side facing away from spring  34  serves as a sealing path for seal  38 . However, that sealing path can also be integrally formed with displaceable disk  35 . That part formed integrally with displaceable disk  35  would then radially inwardly hold spring  34  radially to the inside with its radially outer area. With spring  34  at the inside, that part can form one sealing path radially at the inside and one radially at the outside. 
       FIG. 4  shows a detail of the set of conical disks  33  on the output side, with the displaceable disk  35   a  on the output side being held on the output side shaft  39  so that it is rotationally fixed but axially displaceable by means of the tooth system  40 . The tooth system  40  is in the form of a multiple-tooth spline, i.e., the tooth profile is repeated a plurality of times around the circumference. Also shown is spring  34   a , which, as already described, applies pressure to the output side displaceable disk  35   a  in the direction of the output side fixed disk, which is not shown in  FIG. 4  and is positioned to the left of output side displaceable disk  35   a , as is shown schematically in  FIG. 3 . 
     In the illustrated extreme left position relative to output shaft  39  of output side displaceable disk  35   a , endless torque-transmitting means 2 runs at the greatest possible diameter due to the closest possible wedge-shaped gap between the two conical disks. Since the set of conical disks  33  on the output side is shown here, the variable speed drive unit of the transmission is in the underdrive position, which serves, for example, for starting up. 
     In the end position shown in  FIG. 4 , the displaceable disk  35   a  on the output side is against stop ring  41 . Stop ring  41  is positioned and held in a groove  42  of output side shaft  39 . Because of its elasticity, stop ring  41  prevents an excessively hard end impact of output side displaceable disk  35   a  in the area of the maximum underdrive adjustment position. Stop ring  41  can be made for example of steel wire mesh of braided steel wire, which makes it possible to ensure that it attains the desired working life while having sufficient elasticity. It would also be conceivable to form stop ring  41  using a spring assembly, for example in the form of a Belleville spring assembly. That makes it possible to produce a desired impact behavior, such as, for example, a progressively increasing counterforce, so that output side displaceable disk  35   a  is cushioned relatively gently in the region before its end impact. The function is comparable to that of a stop bumper. 
       FIG. 5  shows an enlarged detail of the area designated V in  FIG. 4 , and illustrates an enlargement of the cross section of stop ring  41 .  FIG. 5  shows, merely as an example, that stop ring  41  can be made of steel wire mesh or braided steel wire  43 , on each axial side of which a bearing shell  44  is positioned. In the example shown, when the steel wire mesh or braided steel wire is compressed the bearing shells  44  form a practically solid stop, because their regions that extend axially toward each other come to rest against each other. 
       FIG. 6  shows, as an example, an end stop in which the damping is accomplished by hydraulic oil. Again, it is the output side shaft  39  and the output side displaceable disk  35   b  that are shown, and again in the region of the maximum underdrive position as in  FIG. 4 . An oil feed conduit  46  branches off from the oil feed conduit  45  shown in  FIG. 6 , which conveys oil to the adjusting unit of the output side displaceable disk  35   b . Oil feed conduit  45  can supply hydraulic medium both to the piston-cylinder unit that is used to adjust the transmission ratio and to the one that is used for clamping. Through the conduit  46  that branches off from oil feed conduit  45  and that can be circumferentially arranged as multiple conduits, hydraulic medium is brought into chamber  47  in the region of the end stop. Chamber  47  is located axially between a shoulder of output side shaft  39  and a corresponding opposing stop of output side displaceable disk  35   b . The hydraulic medium present there is then pressed out or displaced from chamber  47 , which becomes smaller as output side displaceable disk  35   b  approaches its stop, causing the speed of displaceable disk  35   b  to be reduced, so that displaceable disk  35   b  comes to a damped stop. The advantage of that solution is that practically no additional components are necessary, and that the hydraulic medium that is already present in the transmission can be used to damp the end impact. 
     An end stop implemented in accordance with the invention is explained on the output side displaceable disk in the maximum underdrive position merely as an example. But that is by no means to be regarded as a limitation, because all of the stops present in the region of the variable speed drive unit can be implemented in accordance with the invention, or a corresponding stop can be carried out, depending, for example, upon the direction of necessary quick adjustment. 
     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.