CN101006291A - Belt-driven conical-pulley transmission, method for producing it, and motor vehicle having such a transmission - Google Patents

Abstract

The invention relates to an automatic transmission in the form of a spherical disk-shaped enveloping gear system comprising pairs of spherical disks located at the input end and the output end and an enveloping means for transmitting torque, the variator being embodied in a structural space-optimized manner.

Description

Cone-disc wound contact transmission, method of manufacture and motor vehicle having such a transmission

The invention relates to an automatic transmission in the form of a conical-disc winding contact transmission, such as disclosed in DE 10 2004 015 215 and other publications, and to a method for its manufacture and a motor vehicle equipped with it.

In a broader sense, an automatic transmission is a characteristic torque converter (Kennungswandler), whose current gear ratio is automatically adjusted according to the current or expected operating conditions, such as partial load, propulsion and environmental parameters, such as temperature, air pressure , the air humidity changes stepwise or steplessly. It includes characteristic torque converters that are based on electrodynamic, pneumatic, hydraulic, hydrostatic principles or on a combination of these principles.

Automation involves different functions, such as starting, gear ratio selection, and the way of changing the gear ratio in different working states. The changing mode of the gear ratio can be understood as the sequential conversion of each gear, and the jump of the conversion gear. and adjusted speed.

The requirements for comfort, reliability and optional construction costs determine the degree of automation, that is to say how many functions are to be carried out automatically.

Often the driver can manually intervene in the operation of the automation or limit it to a few individual functions.

As used today mainly in motor vehicle construction, automatic transmissions generally have the following structure in the narrow sense:

On the drive side of the transmission there is a starting unit in the form of an adjustable clutch, for example a wet or dry friction clutch, a hydrodynamic clutch or a hydrodynamic torque converter.

For torque converters, a clutch clutch is usually connected in parallel to the pump and turbine section, which increases the efficiency by direct force transmission and damps vibrations at critical speeds by a defined slip.

The starting unit drives a mechanical, continuously variable or stepped transmission, which can include a forward/reverse unit, a main group, a range group, a split group and/or a continuously variable transmission. The gear transmission group is designed as a countershaft or planetary structure with straight teeth or helical teeth, depending on the running smoothness, position conditions and transmission possibilities.

The output member of a mechanical transmission, i.e. a shaft or a gear, drives a differential transmission either directly or indirectly through intermediate shafts or intermediate stages with a constant transmission ratio, which may be constructed as a separate transmission or automatically An integral part of the transmission. The transmission is in principle suitable for longitudinal and transverse installation on a motor vehicle.

Hydrostatic, pneumatic and/or electric actuators can be provided for adjusting the transmission ratio of the mechanical transmission. Hydraulic pumps operating according to the extrusion principle supply pressurized oil for the starting unit, in particular the hydrodynamic unit, for the hydrostatic actuators of the mechanical transmission and for the lubrication and cooling of the system. Depending on the required pressure and delivery volume, gear pumps, screw pumps, vane pumps and piston pumps can be considered, and the latter is usually a radial structure. In practice, both gear pumps and radial piston pumps are used for this purpose, gear pumps being advantageous due to their low construction cost and radial piston pumps due to their high pressure levels and good adjustability.

The hydraulic pump can be located anywhere in the transmission and on a main or countershaft that is continuously driven by the drive unit.

A continuously variable automatic transmission is known, which consists of a starting unit, a planetary reversing transmission (Planetenwendegetriebe) as forward/reverse unit, a hydraulic pump, a continuously variable transmission, an intermediate shaft and a differential . The continuously variable transmission is composed of two cone-disc pairs and a winding contact device. Each cone pair includes a second cone that is movable in the axial direction. Between these pairs of cones runs a winding contact device, which can be, for example, a pusher belt, a traction chain or a belt. The travel radius of the winding contact device and thus the transmission ratio of the continuously variable automatic transmission are changed by the movement of the second conical disk.

Continuously variable automatic transmissions require high pressure levels in order to adjust the conical pulleys of the continuously variable transmission to all operating points with the desired speed and, moreover, to transmit torque as wear-free as possible with sufficient basic contact pressure.

In motor vehicles, high demands are generally placed on comfort, in particular also with respect to acoustics. Especially in high-end motor vehicles, the driver and occupants of the motor vehicle do not wish to have disturbing noises which are generated by the operation of the motor vehicle's unit. However, internal combustion engines and other units such as transmissions produce noises which can largely be perceived as disturbances. For example, in the case of a continuously adjustable transmission, the use of connecting plate chains will cause noise, because this connecting plate chain produces repeated vibrations due to the structure of the connecting plates and pins during the operation of the transmission by the pins hitting the transmission cone. shock. Acoustic effects in CVT transmissions are generally attributed to pin entry ("shock") as an excitation source. This acoustic excitation then resonates at the natural frequency of the transmission housing (FE-mode) or the shaft (torsional mode, bending mode).

Another acoustic effect arises from the CVT chain, which can generate vibrations on the clamping or return section, eg on one side, which can be prevented, for example, by slide rails. Torsional frictional vibrations are known, for example, as tug at a frequency of 10 Hz from a clutch. Tear is excited when the change in the coefficient of friction is a decrease in the coefficient of friction under changing slip conditions. First of all the steel-to-paper friction coefficient is of great relevance in automatic transmissions.

A part of the object underlying the present invention is to improve the acoustics of such a transmission and thus the comfort, in particular the noise-suppression comfort, of a motor vehicle equipped with such a transmission. Another part of the task underlying the present invention is to take appropriate countermeasures based on the analysis of the strong CVT vibrations at high frequencies and the elucidation of the corresponding mechanism of action in relation thereto, in order to minimize or suppress as much as possible the audio frequencies mainly in the order of 400-600 Hz This vibration in the area. A further part of the object of the invention is to increase the operating strength of the components and thus to extend the operating life of such an automatic transmission. A further part of the object of the invention is to increase the torque transmission capacity of such a transmission or to transmit greater forces via components of the transmission. Furthermore - and this is another part of the task - such a transmission should be economically producible.

This object is achieved by the invention as described in the claims and described in the description and abstract with reference to the drawings and further developments thereof.

The analysis yields a realistic understanding of the type of vibration pattern that involves the movement of the chain in winding contact in connection with the tipping and/or bending of the corresponding moving cone, the decisive factor for the vibration frequency is above all the chain The mass and overall rollover and bending stiffness of the moving cone. It will be understood that this stiffness includes the dish structure of the cone itself, tipping of the cone, bending of the shaft due to its elasticity and tilting of the shaft due to different support stiffnesses. In addition, the level and variation of the coefficient of friction as well as the rotational speed and gear ratio are decisive for the frequency.

This recognition is unexpected in this respect, because the vibration of the chain in the winding contact arc - that is, during its clamping in the cone-disk pack - has not been described so far and is in contrast to the hitherto held view that in this arc It is paradoxical that frictional contact with the cone in the segment can damp this vibration.

Also, the influence of CVT oils on such frictional vibrations has not been described so far, so that the oils have hitherto only been improved with regard to a high and temporally stable coefficient of friction and low wear.

While it is known that the tipping clearance between the shaft and the moving cone in a movable CVT cone (moving cone) has some effect on efficiency, no one has so far described the bending, tilting, flipping or swinging motion.

In the case of a CVT transmission in the form of a conical-cone-cone transmission with a conical-cone contact, in particular a chain, the conical-disc transmission of the continuously variable transmission is deformed by the contact force on the conical-cone contact. These contact forces are necessary in order, on the one hand, to prevent slippage of the chain during torque transmission and, on the other hand, to be able to adjust and change the continuously variable transmission and thus the gear ratio of the transmission. Here, the wedge gap formed by the cone halves will change under load. Considering the shape of the cone and the position of the corresponding load point of the winding contact device, when the load generated by the pressing force acting on the winding contact device is the largest and the corresponding force point is placed at the outermost radial direction, it is as large as possible. The wedge-shaped gap deforms maximally from the unloaded position at a diameter of . In a CVT in the form of a conical-disk winding transmission, the point of force application of the winding contact or the chain or the propulsion track will as a rule be determined by the transmission ratio of the continuously variable transmission. Furthermore, it should be taken into account that the point of application of the force does not act on the entire 360° circumference of the conical disk, but only in a certain angular range, which is limited by the corresponding gear ratio and is therefore smaller. This results in an asymmetric disk structure of the cone half, as will be described in detail below.

Due to the non-uniform disk structure and uneven load distribution within the winding contact device, the winding contact device is forced to move radially towards the center of the shaft when the winding contact device runs on the conical disk. The direction of rotation also has an influence on this, since it depends on whether the chain segment in question is part of a loaded segment or an unloaded segment.

The greater the load, the more pronounced this deformation and the greater the friction force and thus the resulting friction travel. This friction leads to loss of efficiency and wear and also acts as an excitation mechanism for frictional vibrations. Frictional vibrations can in turn generate noise, for example, through structure-borne sound excitation.

A case of the aforementioned effect, which is critical to the design, occurs on the driven-side cone at start-up of the cone-wound contactor transmission. The load generated by the drive during start-up is at a maximum and the pressing force on the winding contact is also at a maximum through the corresponding gear ratio for shifting to slow speed. Due to the gear ratio, the winding contact or the chain is located radially outermost on the output side cone. Due to this load, the driven-side cone is deformed or pushed apart so strongly that the wedge gap becomes very large, resulting in a maximum friction stroke and friction force.

For solving this task and improving transmissions according to the prior art in which, for example, four conical disks are geometrically similar in terms of disk shape and rigidity, a conical-disc winding contact device according to the invention Contributed by a transmission, it has drive-side and driven-side cone pairs, each with a fixed cone and a movable cone, which are each arranged on a drive-side and a driven-side shaft and can be The connection is via a winding contact, wherein the cone-disk winding transmission has a continuously variable transmission optimized in terms of stiffness.

Furthermore, contributing to the solution of this task and the improvement of transmissions according to the prior art is a conical-to-cone contact transmission having drive-side and driven-side conical-disk pairs, each of which has a fixed cone and a Moving cones, which are each arranged on a drive-side and a driven-side shaft and can be connected by a winding contact device for torque transmission, wherein a sliding fit seat of at least one moving cone is arranged on In the radially inner region of the moving bevel and at least one sliding fit seat of at least one moving bevel is arranged in the radially outer region of the moving bevel.

In a sliding fit seat arrangement close to the shaft, as shown for example in FIG. 1 and in FIGS. Cone length has an effect. If one of these sliding fit seats is moved to the radially outer side, the following parts can be arranged under this sliding fit seat, so that these parts can be located radially inside the radially outer sliding fit seat, Axial installation space can thus be saved. In this space radially inside the sliding seat, for example, a bearing for the cone-disk pack, a housing part with a rotary introduction for supplying fluid to the respective cone-disk pack, a hydraulic pump or a drive unit for the hydraulic pump can be arranged .

In addition, it is also possible, for example, to use the additional new installation space obtained in the interior region for the distribution gear of the all-wheel arrangement.

It may be particularly advantageous if, in the conical-disc winding contact gear transmission according to the invention, the movable cone has two sliding seats, whereas it may be advantageous, for example, in terms of rigidity, that the movable bevel has three sliding seats, For example as shown in FIG. 10 and described in this regard.

Furthermore, it can be advantageous to form a centrifugal oil bellows using radially outer sliding seats, whereby for example additional installation space is available in the radially inner region.

In a conical-pulley wound transmission according to the invention, the sliding fit seat arrangement can be arranged on the drive-side and/or output-side conical-pulley pair.

Since the necessary additional axial installation space length of the sliding seat due to a seal arranged on the extension of the sliding seat is not decisive for the installation space, the radially outer sliding seat can be arranged by means of an axial It seals upwardly with the sealing device arranged adjacent to it.

It can generally be advantageous in a conical-disc wound contact transmission according to the invention that the bearing of the movable conical disc is arranged radially inward of the sliding fit seat arranged radially outwardly.

For example, it may be advantageous in terms of a configuration suitable for machining that the radially outer sliding seat is formed here using a part which is connected to the moving cone, wherein the connection can be a welded connection .

In addition, when using this component here, a centrifugal oil cover can be formed which can be used for speed-dependent centrifugal oil compensation, wherein two centrifugal oil chambers can also be formed in order to achieve a higher centrifugal oil compensation.

It can be particularly advantageous if the stiffness is designed to be significantly greater in the output-side cone set than in the drive-side cone set when forces act radially outward. In this case it has proven to be advantageous if the stiffness is made 1.2 to 1.6 times greater.

It can also be advantageous if the output-side rotary cone has a significantly higher stiffness than the drive-side rotary cone.

In an inventive conical-wound conical transmission, it may be advantageous if the output-side conical disk has a geometrically significantly stronger conical disk than the drive-side conical disk.

Furthermore, it may be desirable for the output-side rotary cone to have a geometrically significantly stronger cone neck than the drive-side rotary cone.

It may prove to be advantageous if the output-side rotor cone has a geometrically substantially stronger cone plate than the output-side fixed cone.

It can be advantageous if the drive-side rotary cone has a geometrically substantially stronger cone disk than the drive-side fixed cone.

It may prove to be desirable if the output-side rotary cone has a smaller mean guide play than the drive-side rotary cone.

Furthermore, it can be advantageous if the output-side rotary cone has a large guide seat which is significantly longer than the drive-side rotary cone.

It may be expedient for at least one movable cone to have at least one seal track formed integrally therewith.

It can also be advantageous if at least one movable cone has two directly connected seal slide rails.

Depending on the embodiment, it may be expedient to produce the seal slide rail by turning or non-machining.

Furthermore, in the adjacent state, apart from the at least one sealing point, there is a free area which can be used as a dirt space.

It may be advantageous in the case of the conical-conical winding contact transmission according to the invention that the driven conical disk on the output side has a cylindrical conical disk neck, wherein the conical disk neck can be used for spring centering, and/or the conical disk neck Has a semicircular groove that can be used for spring contact.

It can generally be advantageous if the output-side thrust cone has a radially far outer pressure spring.

Furthermore, it can be advantageous if the output-side rotor cone has at least one laid-on sheet metal part which can be used as a seal runner for the at least one seal.

Depending, for example, on the design of the continuously variable transmission, the spring can be cylindrical, waisted or conical.

It can often be advantageous if the driven-side fixed disk has a significantly higher stiffness than the driven-side fixed disk.

It may be particularly advantageous to construct the continuously variable transmission according to the double-piston principle, as described, for example, in DE 103 54 720.7.

Solving the problem may require the consideration of more than one influenceable parameter and, for example, the combination of certain properties of the oil with a certain machine configuration.

In order to solve this task, a conical-disc winding contact transmission can also be proposed, which has a drive-side and a driven-side conical-disc pair, each of which has a fixed conical disc and a movable conical disc, which are each arranged in a The drive side and a driven side are connected on the shaft and can be connected via a wound contact device for torque transmission, wherein at least one of the factors listed below is optimized in terms of transmission acoustics:

- viscous or hydraulic media in the form of oil,

- the surface properties of the contact area between the cone-disk and the wound contact device,

- the geometry of at least one cone,

- damping of at least one cone,

- Guiding of at least one cone.

In this case it may be advantageous to use an oil with a coefficient of friction that is insensitive to the friction speed. Furthermore, it may be advantageous to optimize the contact surface between the conical disk and the wound contact device, for example with respect to its topography.

Furthermore, it may be advantageous if at least one rigidity-optimized conical disk and/or at least one damped conical disk is provided. It has also proven to be advantageous to combine at least one radially outer conical disk in the transmission.

The invention also relates to a method for producing a transmission according to the invention.

Furthermore, the invention relates to a motor vehicle having a transmission according to the invention.

The invention will be described in more detail below by way of example with the aid of the schematically represented drawing.

The accompanying drawings indicate:

Figure 1: Partial view of a cone-disk wound contact device transmission,

FIG. 2 : a view substantially corresponding to FIG. 1 of another embodiment,

Figures 3 and 4: Graphs of friction coefficient relationships,

Figures 5 and 6: Schematic configuration possibilities of the mantle,

Figure 7: Schematic diagram of a disc structure with cone-disk asymmetry,

Figure 8a: Cone-disk wound contact device transmission with geometrically similar cone-disk packs,

Figure 8b: Cone-disk wound contact transmission with stiffness-optimized cone-disc pack,

Figures 9 and 10: Example of a pair of driven side cones,

Figures 11 and 12: A cone-disk stack on the drive side,

Figure 13: An enlarged view of area XIII in Figure 11,

Figure 14: A cone-disk pack on the driven side,

Figure 15: A partial view of the driven-side cone-disk pack.

FIG. 1 shows only a part of a conical-cone transmission, ie the drive-side or input-side part of the conical-cone transmission 1 , which is driven by a drive motor, for example an internal combustion engine. In a fully formed conical-cone-conical transmission, a complementary formed output-side part of a continuously adjustable conical-conical transmission is assigned to the input-side part, wherein the two parts are connected via an e.g. Wound contacts in the form of connecting leaf chains 2 are connected to each other for torque transmission. On the input side, the conical-wound contact transmission 1 has a shaft 3 which, in the exemplary embodiment shown, is formed in one piece with a fixed conical or fixed conical disk 4 . The axially fixed cone 4 is located adjacently opposite an axially movable cone or moving cone 5 in the axial direction of the shaft 3 .

In the view according to FIG. 1, the connecting plate chain 2 is shown at a radially outer position on the pair of drive-side cones 4, 5, which is obtained in that the axially displaceable cones in this figure The movement of the disc 5 in the right direction and this axial displacement movement of the axially displaceable conical disc 5 causes the connecting leaf chain 2 to move in the radially outward direction, thereby changing the gear ratio of the transmission to fast.

The axially displaceable conical disk 5 can also be moved to the left in the drawing in a manner known per se, wherein in this position the connecting leaf chain 2 is located in a radially inner position (indicated by reference numeral 2a), in which the conical disk The gear ratio of the winding contactor transmission 1 is changed to a slow speed.

The torque provided by a drive motor, not shown, is introduced into the drive side part of the conical-disc winding contact transmission shown in FIG. A rolling bearing in the form of a ball bearing 7 for axial force is supported on the shaft 3 , which ball bearing is fixed on the shaft 3 via a retaining ring 8 and a shaft nut 9 . A torque sensor 10 is arranged between the gear 6 and the axially movable cone 5, and an axially fixed support plate 11 and an axially movable support plate 12 are arranged on the torque sensor. Support plate configuration 13. Rolling bodies, for example in the form of balls 14 shown, are arranged between the two support discs 11 , 12 .

The torque introduced by the gear 6 leads to an angle of rotation between the axially fixed support disc 11 and the axially movable support disc 12, which will cause an axial displacement of the support disc 12, precisely due to Start ramps are provided on the support discs, on which the balls 14 move upward and thus serve for the axial offset of the support discs relative to each other.

The torque sensor 10 has two pressure chambers 15 , 16 , wherein the first pressure chamber 15 is provided for being acted upon with pressure medium depending on the introduced torque and the second pressure chamber 16 is supplied with pressure medium, precisely according to the transmission. The transmission ratio is supplied to the pressure medium.

A piston/cylinder unit 17 is provided for generating a pressing force, by means of which a normal force is applied to the connecting plate chain 2 between the axially fixed conical disk 4 and the axially movable conical disk 5, which The piston/cylinder unit has two pressure chambers 18 , 19 . A first pressure chamber 18 is used to vary the loading of the link chain 2 as a function of the transmission ratio, and a second pressure chamber 19 is used to increase or decrease the contact force in connection with the torque-controlled pressure chamber 15 of the torque sensor 10, The connecting leaf chain 2 between the cone pairs 4, 5 is acted upon by this pressing force.

To supply these pressure chambers with pressure medium, the shaft 3 is provided with three channels 20 through which pressure medium from a pump (not shown) is supplied into the pressure chambers. Via a discharge-side channel 21 pressure medium can flow out of the shaft 3 and be fed back into the circuit.

The loading of the pressure chambers 15 , 16 , 18 , 19 leads to a torque- and transmission-ratio-dependent displacement of the axially displaceable conical disk 5 on the shaft 3 . In order to receive the movable cone 5 , the shaft 3 has centering surfaces 22 which serve as sliding fit seats for the movable cone 5 .

As can be easily seen from FIG. 1 , the conical-wound contact transmission 1 each has a noise damping device 23 in the region of the bearing points of the conical disks 5 on the shaft 3 . For this purpose, the noise damping device can have an annular body and a damping lining or consist of only one damping lining.

The reference numbers used in FIG. 1 also refer to comparable features in the other figures. In this respect these drawings should be considered as a unit. For the sake of clarity, only reference numbers other than those in FIG. 1 are used in the other figures.

In Fig. 2, the central channel of the three channels 20 is formed in a modified form relative to Fig. 1; it can be seen that the hole 24 forming the central channel 20 is made significantly shorter than the central hole in Fig. 1, the hole 24 is machined as a blind hole from the side shown on the right in FIGS. 1 and 2 . Such blind holes are expensive to manufacture and require high precision during processing. In this case, the manufacturing costs and the demands on the reliability of the machining process increase in proportion to the length. Such a shortening of the bore therefore has an advantageous effect on the production costs.

A transverse bore 25 branches off in the bottom region of the bore 24 , wherein a plurality of transverse bores can also be arranged distributed over the circumference. In the illustrated case the transverse bore 25 is designed as a radial bore, but it can also be formed at another angle as an oblique bore. The bore 25 penetrates the surface of the shaft 3 at a point which, irrespective of the operating state, ie, for example, regardless of the set gear ratio, is located in a region which is always covered by the passive conical disk 5 .

By arranging the transverse bore 25 in the area covered by the movable cone 5, the shaft 3 can be made shorter in the axial direction, thereby saving installation space. In addition, a reduction in the load can also be achieved by shortening the shaft 3 .

Here, the outlet of the channel or transverse bore 25 can be arranged, for example, in the region of the turned groove 26 adjacent to the centering surface 22 of the shaft. This is particularly advantageous if the toothing 27 , which connects the movable cone 5 axially displaceably but non-rotatably to the shaft 3 , is subjected to high loads, for example due to transmitted torque.

In many cases, however, the loading of the tooth structure 27 is not a critical design criterion, whereby the outlet of the hole 25 can be provided in the region of this tooth structure, as shown in FIG. 2 . Arranging the transverse bores 25 in the toothing 27 instead of in the turned grooves 26 has the advantage that there is a greater resistance torque through which the bending stresses in the edge fibers are reduced. Furthermore, if the critical fibers disturbed by the transverse holes 25 are kept at approximately the same radius, the moment of inertia of area is greater at this location. This results in a significant reduction in stresses in the critical area around the outlet of the transverse bore 25 between the teeth of the tooth structure 27 . The supply of hydraulic fluid is the same in FIGS. 1 and 2 because the pressure chambers 15 and 19 are connected to each other and the moving cone 5 has connection holes 28 which connect the area of the tooth structure 27 to the pressure chamber 19 . In these figures the mantle 5 is shown in its outermost left position, which corresponds to the starting gear ratio or underdrive. If the movable cone 5 is now moved to the right in the direction of the fixed cone 4, a part of the cavity or chamber 29 is always located on the outlet of the transverse hole or channel 25, so that the required fluid supply. Also as in FIG. 1 , there are two switching states for the pressure chamber 16 , which are dependent on the axial position of the movable cone 5 . In the position shown, the control hole 30 is exposed, thereby depressurizing the channel 20 connected to it and closed axially by a plug 31 and the pressure chamber 16 connected to it by a channel not shown or only with environmental stress. If the movable cone 5 is now moved in the direction of the fixed cone 4, it will move over these control holes 30, wherein the chamber 29 is located above the outlet of the control holes 30 from a certain stroke. However, there is a high torque-dependent high pressure in the chamber 29 which then also passes through the control bore 30 and the channel 20 into the pressure chamber 16 , so that a high pressure is also applied there. In this way, two switching states are realized, which control the pressing force as a function of the transmission ratio.

In addition, a disk spring 32 is provided in Fig. 2, which moves the movable cone 5 to a predetermined axial position in the depressurized state of the transmission 1, thereby adjusting a gear ratio of the transmission 1, the gear ratio Prevents excessive loads, e.g. when a motor vehicle is towed.

FIG. 3 shows two diagrams which show the curve of the coefficient of friction versus the sliding speed as a function of the contact pressure. Here, the sliding velocity is shown on the abscissa and the friction coefficient on the ordinate. The dashed line is to be regarded as a reference value and represents a coefficient of friction, which can lie, for example, at approximately μ=0.12. As can be seen from these two figures, the coefficient of friction is a function of sliding speed, where the coefficient of friction tends to decrease with increasing sliding speed.

As already mentioned above, for example in clutches the coefficient of friction decreases with increasing slip speed, leading to tearing and thus to a reduction in comfort. It is therefore attempted to keep the drop in the coefficient of friction over speed as small as possible.

The curve shown in FIG. 3 occurs at the points of contact between the pendulum pressure elements of the chain and the contact surfaces of the cones that interact with them. In this case, the chain or the wrapping contact device is acted upon both in the direction of travel by the torque to be transmitted and in a direction transverse to the direction of travel by predominantly pressing forces. Here, the contact force should be selected in such a way that the torque to be transmitted can be transmitted to the other cone-disc set with sufficient reliability and without slippage.

The corresponding distance of these curves in the direction of the ordinate represents the spread width of the coefficient of friction as a function of contact pressure or contact pressure. Here the lower line corresponds to a low contact pressure and the upper line corresponds to a high contact pressure respectively.

A comparison of the previously known embodiments according to the upper graph with the embodiment of the invention according to the lower graph clearly shows that firstly the dispersive areas each delimited by the two curves become smaller, resulting in the coefficient of friction There is less dependence on the contact pressure or pressing force just applied. In other words, the embodiment according to the invention (lower graph) is insensitive to changes in the contact pressure.

It can also be seen from FIG. 3 that the curve in the lower graph runs flatter, which results in a smaller dependence of the friction coefficient on the sliding speed. A more stable behavior of the friction coefficient can be achieved by this flat negative gradient of the friction coefficient with respect to the sliding velocity. In this case it is less of a problem when the curves move approximately parallel from top to bottom or vice versa than when they change in their slope, since any change in slope represents a greater dependence of the coefficient of friction on the sliding velocity.

The curves of the coefficient of friction versus sliding velocity and relative contact pressure thus clearly determined, as shown in the lower graph of Figure 3, can result in damping of vibrations that occur through the steel-steel contact between the belt or chain and the cone. The change in the coefficient of friction of the contact is excited. Vibrations can be overcome directly at the point where they occur by means of such a change in the coefficient of friction by using a corresponding oil.

The graph in FIG. 4 is created essentially like the graph in FIG. 3 , but instead of showing the dependence on the oil used, the dependence on surface characteristic values is shown instead. The explanations made in connection with FIG. 3 and the remarks about the improvement also apply to FIG. 4 , ie the lower graph shows a significant improvement in the situation.

The upper graph of FIG. 4 shows the situation on a polished surface, while the lower graph of this figure shows the dependence of the coefficient of friction on sliding velocity and contact pressure for the surface characteristic values according to the invention. These surface characteristic values can be produced, for example, by a polishing process in which the friction parameters have a correct profile and can be maintained over long operating times. For example, noise phenomena immediately appear on smooth surfaces, whereas in the most favorable case no noise occurs at all in the state of rough surfaces. This improvement in the noise behavior can also be achieved by a reduction in contact force or contact pressure.

Tests with the aid of simulation calculations and measurements have shown that the vibration and noise behavior can be positively influenced by an increase in the tilting rigidity of the axially movable cone or moving cone, where this is especially - but not exclusively - suitable for The driven cone on the driven side. Overall, the increased flexural strength, by means of which the opening of the conical disks, in particular the driven-side conical disk pack, can be reduced reduces the vibration amplitudes which are relevant for the noise generation. A similar effect can be achieved by increasing the damping at this location.

Figures 5 and 6 each schematically show the profile of a moving cone, wherein each figure only shows the upper half of the rotationally symmetrical profile.

FIG. 5 shows the reinforcement structure of the conical disk itself in schematic embodiments a) to e). FIGS. 5 and 6 here each schematically show a portion of the driven-side, axially displaceable cone or rotor 33 , wherein a similar configuration can also be transferred to the drive-side rotor 5 .

The moving cone 33 shown in FIG. 5 a has in its region facing away from the winding contact device 2 a plurality of reinforcing ribs 34 arranged distributed over the circumference, which reduce - or in the most favorable case prevent - The expansion of the radially outwardly projecting part of the cone 33 under load can thus counteract the expansion of the cone pair.

The rotor 33 according to FIG. 5 b has a configuration in which the radially outwardly projecting region of the rotor 33 is reinforced in such a way that its wall thickness increases radially outward. This is achieved by a corresponding configuration of the contour of the cone facing away from the winding contact device 2 . The continuous change of the contour shown here can also be modified in that the wall thickness increases in steps.

In order to reinforce the movable cone 33 in the axial direction, a reinforcing collar 35 can also be provided radially outside, as shown in FIG. 5c. Figure 5d shows that, in addition to the radially outer reinforcing collar 35, a further reinforcing collar 36 is provided, which is arranged radially inward and can thus also be used as between two pressure chambers if required spacer between.

In FIGS. 5c and 5d reinforcing collars 35 and 36 are shown as separate parts or rings which have to be connected to the moving cone 33 . FIG. 5e shows a possibility of forming the reinforcing collar 35 and/or the reinforcing collar 36 integrally with the rotary cone, wherein a machining-compatible design can advantageously be taken into account.

Figures 5f and 5g illustrate the reinforcement of the connection of the cone to the shaft. On the one hand, the hub 37 of the movable cone 33 is connected with the radially outwardly extending part of the movable cone 33 through a reinforcement ring 38, so that at least the deformation of this area can be reduced, and reinforcement ribs 34 are also provided. , these reinforcing ribs are connected to the reinforcing ring 38 on the one hand and to the hub 37 of the moving cone 33 on the other hand.

FIGS. 6a to 6e show the basic damping possibilities for an axially movable cone or rotor 33 on the output side, but they can also be used for an axially movable cone or rotor 5 on the drive side. .

FIG. 6 a firstly shows that the hub 37 is divided into individual layers, the stack of layers being compressed by the compression pressure exerted by the hydraulic medium and thus having a damping effect.

In FIG. 6 b additionally the reinforcing collar 35 is formed as a stack of layers, which in turn is compressed by means of a pressing pressure. According to FIG. 6 c , the radially inner reinforcing collar 36 can also be configured as a stack, wherein the reinforcing collar 36 can also be used as a separation structure between different pressure chambers. Alternatively, in an embodiment according to FIG. 6c the hub 37 can also be divided into individual layers.

FIGS. 6 d and 6 e each show springs 39 , which increase the friction between the cylinders of the individual layers through the additional radial compression, thereby simultaneously increasing the damping effect. In FIG. 6e the hub 37 can also be configured as a stack.

Another solution is shown in FIGS. 6f and 6g, which consists in changing the tilting direction of the movable cone. The radially outer region of the rotor exhibits the greatest deflection in the tilting direction when the rotor is generally guided via its radially inner region or via its hub 37 . In order to counteract this problem, the mantle can in principle be guided on the outside, so that the mantle rests with its radially outer region on the outer guide 40 and thus cannot deviate there. The deflection can then occur in the radially inner region of the movable cone 33 , for which the measures described can also be taken. In this case, however, care should be taken to avoid tilting or clamping of the movable cone 33 between these guides.

FIG. 7 schematically shows the driven-side rotor 33 , a similar effect also occurring at the drive-side rotor 5 . The statements made for the driven-side rotary cone 33 therefore also apply to the drive-side rotary cone 5 ; for the sake of clarity, only the processes and features are described below with reference to the rotary cone 33 .

The conical disk 33 consists of two main areas, namely the conical disk 42 and the conical neck or hub 37 . The movable cone 33 is supported in a non-rotatable but axially displaceable manner on the output-side shaft 41 and thus transmits the torque introduced by the winding contact device 2 to the output, that is, for example via a differential The transmission device and the drive shaft connected to it through the flange are finally transmitted to the drive wheels of the motor vehicle.

7 shows not to scale the two contours of the moving cone 33, namely contour A representing the undeformed, unloaded state and on the other hand contour B representing the deformation obtained under the action of force F state. In this regard, it can be determined that the unloaded, undeformed state according to the contour A is rotationally symmetrical as indicated by the figure.

The force F indicated by the radially outer arrows is the reaction force of the wrap-around contact to the sum of the aforementioned pressing forces for transmission torque transmission and gear ratio adjustment. On the point of action of the force F shown and along an arc section that extends over some parts of the circumference of the movable cone 33, it touches the winding contact device 2, while winding on the opposite side (the side shown below) The contact device 2 is not in contact with the moving cone 33 because the winding contact device 2 extends in the direction towards the complementary cone group.

As can be seen from FIG. 7 , the profile change from profile A to profile B results not only from the deformation of the conical disk 42 but also from the tilting of the entire conical disk 33 . If only a deformation of the conical disk 42 had occurred, the contours A and B would be identical on the unloaded side shown below.

However, it can be seen from the figure that the deformed profile B on the loaded side is deflected in the direction of the force F acting on it (to the right in the figure), while on the unloaded side the profile B is deflected in the same direction as Force F is deflected in the opposite direction (to the left in FIG. 7 ).

This deflection is caused by the tilting of the entire movable cone 33, because on the one hand the cone neck or disk hub 37 also has only a limited rigidity and on the other hand due to the axial displaceability of the cone or the movable cone 33 the cone The disc cannot be guided over its entire length cooperating with the shaft 41 . In addition, a certain guide play between the disk hub 37 and the shaft 41 is required for the axial mobility, but on the other hand this promotes tilting of the movable cone 33 . The larger the gap is made in this case, the more noticeable the tipping will be.

The bending moment generated by the force F relative to the corresponding conical disk circumference produces both deformation and tipping, which bending moment (at a constant force) increases with increasing radius over which the winding contact device 2 travels.

Due to the tilting and uneven deformation of the cone 33 and the uneven load distribution inside the winding contact device 2, the winding contact device 2 will be forced to carry out a radial movement when it is wound on the cone, wherein the The chain or winding contact device 2 moves radially inwards in the direction of the shaft 41 . As a function of load and deformation, the friction force and thus the resulting friction path increase strongly. This results in poorer efficiency and increased wear on the interacting surfaces. It has also been determined that this is the excitation mechanism for frictional vibrations which in turn generate noise through structure-borne sound excitation.

8a and 8b each show a continuously variable transmission 43 with a drive-side conical-disc set 44 and a driven-side conical-disc set 45, wherein in FIG. Stage transmission 43.

The drive side cone group 44 always has a fixed cone 4 and a moving cone 5, which are connected to the corresponding moving cone 33 and the fixed cone of the driven side cone group 45 through a winding contact device in the form of a connecting plate chain 2. The cones 46 are connected.

The reference numerals 47 to 56 used in Figures 8a and 8b have the following meanings;

47-Neck diameter of driving cone

48-Neck diameter of driven side cone

49-Width of driving side conical disc

50-drive side fixed cone disc width

51-Width of driven side fixed cone

52-Width of driven side conical disk

53-Length of small sliding fit seat on driving side

54-Large sliding fit seat length on drive side

55-The length of the large sliding fit seat on the driven side

56-The length of the small sliding fit seat on the driven side

In the continuously variable transmission 43 according to FIG. 8 a , the drive-side and output-side thrust cone diameters 47 and 48 are virtually identical, ie have comparable diameters and thus comparable strengths. In addition, it should be confirmed that the width of the movable cone and the width of the fixed cone 49, 50, 51 and 52 on the driving side and the driven side have roughly comparable results, so that the corresponding cones 4, 5, 33 and 46 The geometries and thus their strengths are of comparable order of magnitude. And the large and small sliding fit seats 53, 54, 55 and 56 on the drive side and driven side are configured to be similarly long in their length, so that there is a geometrically comparable relationship in this regard, especially its Involves the support of the respective mantle on the shaft to which it is attached.

The rigidity-optimized continuously variable transmission 43 according to FIG. 8 b is designed differently. The driven cone diameter 48 is designed significantly larger than the driven cone diameter 47, wherein the driven cone diameter is simultaneously configured as a guide diameter for the pressure spring 57 arranged thereon . The compression spring 57 shown in FIG. 8 b is cylindrical, whereas according to FIG. 8 a it can also be designed to be waisted. A conical configuration of the compression spring 57 is also possible.

An increased rigidity of the driven disc 33 is achieved by the enlarged driven disc neck diameter 48 because an increased bending resistance moment is thereby achieved.

Furthermore, it can be seen from the illustration according to FIG. 8 b that the driven-side conical disk set 45 is designed to be significantly more rigid than the drive-side conical disk set 44 . The comparison shows that the width 51 of the driven side fixed cone is greater than the width 50 of the driven side fixed cone. Furthermore, the driven-side conical disk width 52 is significantly greater than the drive-side conical disk width 49 . And the lengths 55 and 56 of the sliding fitting seats on the corresponding driven sides of the large and small sliding fitting seats are greatly greater than the lengths of the corresponding sliding fitting seats of the drive side cone-disc group 44 , and these lengths are provided with reference numerals 53 and 54 .

This results in an increased stiffness of the driven-side conical disk group 45 compared to the drive-side conical disk group 44 , on the one hand due to the increased strength of the conical disks 33 and 46 due to their increased dimensions. On the other hand, due to the improved support through the increased sliding fit lengths 55 and 56 , better protection against tipping under the load of the traction mechanism 2 is obtained.

In order to further increase the rigidity against tilting, the clearance can be reduced, through which the movable cone 33 is mounted axially displaceable but non-rotatably on sliding fit seats 55, 56 on the shaft, so as to thereby also resist The effect of moving cone 33 tipping tendency.

In summary, the following configurations play a role in optimizing the stiffness of the continuously variable transmission 43:

- the driven-side cone set 45 is reinforced by the geometry of the cones 33 and 46 relative to the drive-side set 44,

- the movable cones 33 and 5 are reinforced relative to the fixed cones 4 and 46,

-The lengths 55 and 56 of the driven side sliding fit seat are increased relative to the length 54 and 53 of the drive value sliding fit seat,

- The driven side cone diameter 48 is enlarged compared to the drive side cone diameter 47,

The large sliding seat 55 on the driven side of the driven cone 33 is designed such that it has as large a guide length as possible in the low gear position (the winding contact runs radially outward).

In principle, although it is possible to make a corresponding modification to the entire continuously variable transmission 43, that is, to provide a stronger cone and an increased length of the sliding fit seat, etc., but for example, it is limited by the structural space and weight that the transmission can provide.

FIG. 9 shows two configuration possibilities of the driven-side conical-disk group 45 , wherein a conical-disk group constructed according to the single-piston principle is shown in the lower part, and a conical-disk group constructed according to the double-piston principle is shown in the upper part. Cone-disk group, for example described by DE10354720.7.

In the double-piston principle, separate pistons are provided for the pressing force and adjustment, while in the single-piston principle there is only one piston/cylinder unit, which introduces the corresponding force into the cone-disc group.

The basic structure of the cone-disk set 45 according to FIG. 9 is as described so far, in particular in connection with FIG. 8 b . What has been said above also applies to the configuration with regard to strength optimization.

Compression spring 57 here has a larger diameter than in the previously described embodiments, wherein its point of action on movable cone 33 is located radially further on the outside. This arrangement also has the advantage that more installation space is available for thickening or geometrically strengthening and increasing the diameter of the conical disk neck or hub 37 . The resulting strength gain has been described above. This results in a variant of the compression spring 57 in the dual-piston principle shown in the upper part of FIG. 9 in that it moves from the radially inner pressure chamber to the radially outer pressure chamber. The sheet metal part 58 which supports the compression spring 57 radially on the inside is fixedly connected to the movable cone 33 and uses its surface facing away from the spring 57 as a seal slide rail for the seal 59 . However, the seal slide rail can also be formed in one piece with the movable cone 33 as shown in connection with FIG. 8 . This part, which is formed integrally with the movable cone 33 , then again holds the compression spring 57 radially inwardly with its radially outer region. With compression spring 57 located on the inside, these components each form a seal slide rail radially on the inside and on the outside.

FIG. 10 shows a further embodiment possibility of the driven-side conical disk pack 45 , for which the previous descriptions also apply, in particular with regard to optimization of its stiffness. As mentioned above, the driven cone 33 is firstly supported on the shaft 41 through the two sliding fit seats 55 and 56 . Compared to the embodiments shown so far, the centrifugal oil bellows 60 are clearly reinforced and made more rigid, so that the actuating cone 33 is additionally supported on the flange part 61 via a sliding fit 62 . If sealing is required in the region of this sliding fit seat 62, this can be achieved by means of a sealing device 63 (top in FIG. 10). The mantle 33 thus has three sliding fits 55 , 56 and 62 by means of which the mantle is supported relative to the shaft. Such a support has a further increased stiffness, whereby this embodiment also contributes to solving the problem on which the present invention is based.

FIG. 11 schematically shows a drive-side cone-disk pack 44 with a starting unit 64 schematically indicated by a dotted line, a torque sensor 10 and a winding contact device in the form of a connecting leaf chain 2 . Here, the radial position of the connecting plate chain 2 depends on the size of the wedge gap, which is increased or reduced between the fixed cone 4 and the movable cone 5 depending on the transmission ratio in such a way that the movable cone 5 moves axially away from or towards the fixed cone 4. Here, in the upper part, a position of the mantle 5 is shown which produces the largest possible transmission ratio of the transmission to underdrive (underdrive). For this reason, the distance between the fixed cone 4 and the movable cone 5 is the largest, that is to say, the movable cone 5 is located at the leftmost position in its figure 11 . In contrast, the lower part shows the maximum gear ratio for fast changeover (overdrive), where the distance between the fixed cone 4 and the movable cone 5 is at a minimum, so that the connecting leaf chain 2 is as large as possible run on the diameter. The mantle 5 is shown in its rightmost position for this purpose.

The movable cone 5 is mounted in a non-rotatable manner relative to the fixed cone 4 but is axially movable. This mounting is achieved on the one hand by the tooth structure 27 and on the other hand by two sliding fit seats 65 and 66 , wherein the first sliding fit seat 65 is arranged radially inside and the second sliding fit seat 66 is arranged on the bearing 67 In the radially outer area of the moving cone 5 outside the radial direction.

A comparison, in particular with FIG. 8 a , shows that due to the radially outer arrangement of the second sliding seat 66 , as shown in FIG. 11 , its radially inner and thus overall axial installation space is saved. For example, a part of the housing basic structure 68 can be arranged in this installation space, in which ducts 20 can be arranged, for example, for the fluid supply of the gear ratio-dependent adjustment of the cone-disk set 44 . A further advantage of the radially outer arrangement of the second sliding seat 66 is that the movable cone 5 can be supported better against tilting, thereby increasing the stiffness of the cone group and avoiding or at least reducing the possibility of This creates disadvantages, as already described above.

FIG. 12 schematically shows that in the radially inner region of the sliding fit seat 66 and the bearing 67 a hydraulic pump 69 indicated by a dotted line can be arranged. The hydraulic pump 69 is used again to provide hydraulic medium under pressure for adjustment and compaction of the cone-disc group. For this purpose, the hydraulic pump 69 is driven via a drive shaft 69 a which is itself driven in the region of the starting unit 64 and is arranged coaxially in the shaft 3 of the cone-disc group 44 .

FIG. 13 shows a detail of XIII in FIG. 11 in an enlarged view. As can be seen from FIGS. 11 to 13 together, the length of the sliding fit seat 66 is not decisive in terms of installation space due to its radially outer arrangement, so that despite the large bearing length of the sliding fit seat 66, However, it is still possible to arrange the sealing device 70 adjacent to the actual sliding fitting seat 66 in the axial direction or to arrange it on the axial extension of the sliding fitting seat 66 without making the length of the sliding fitting seat 66 decisive. shorten. The relatively large length of the sliding seat 66 in turn has a favorable effect on the rigidity of the rotary cone and thus the entire continuously variable transmission, for example. The sealing device 70 is needed because on the one hand the sliding fit seat 66 must have a certain clearance to ensure axial movability, and on the other hand there is hydraulic pressure on the side of the sliding fitting seat 66 facing away from the sealing device 70 , which The hydraulic pressure comes from the adjustment and compression of the cone, while the side of the sealing device 70 facing away from the sliding fit seat 66 has practically ambient pressure, resulting in a strong pressure drop.

FIG. 14 shows a driven-side cone set 45 which in turn has a radially inner sliding seat 65 and a second radially outer sliding seat 66 . In this case, the second sliding fit 66 is formed using a centrifugal oil cover 60 which is supported on the one side by means of the sliding fit 66 on the base structure and on the other side by means of a weld seam. 71 is connected with driven side moving cone 33. The oil located in the centrifugal oil chamber 72 brings about a speed-dependent centrifugal oil compensation. In the radially inner region of the sliding fit seat 66 , which is formed by the displacement of the sliding fit seat 66 towards the radially outer side, for example, a distributing gear 73 of an all-wheel arrangement can be arranged, which is shown in FIG. 14 by Dotted lines schematically indicate. The torque introduced into distribution gear 73 is distributed via it to two output shafts, one of which drives the front wheels of the motor vehicle and the other the rear wheels of the motor vehicle, for example.

The embodiment shown in FIG. 15 substantially corresponds to the embodiment according to FIG. 14 , wherein here a further centrifugal oil chamber 74 is formed in addition to the centrifugal oil chamber 72 for further speed-dependent centrifugal oil compensation.

List of reference signs

1 Cone-disc winding contact device transmission 2a The radial inner position of the connecting plate chain

2 connecting leaf chains 3 axes

4 fixed cones 31 plugs

5 moving cones 32 disc springs

6 gears 33 moving cone (driven side)

7 Ball bearings 34 Ribs

8 retaining ring 35 reinforced convex ring (outer)

9-axis nut 36 reinforced convex ring (inner)

10 torque sensor 37 hub

11 Axially fixed support plate 38 Reinforcing ring

12 axially movable support plate 39 spring

13 support plate configuration 40 outer guide

14 balls 41 axis (driven side)

15 first pressure chamber 42 conical disk

16 Second pressure chamber 43 Continuously variable transmission

17 piston-/cylinder unit 44 drive side (cone) disc group

18 The first pressure chamber 45 Driven side (cone) disc group

19 Second pressure chamber 46 Fixed cone driven side

20(3) channels (feed) 47 mantle neck diameter drive side

21 channels (exit side) 48 moving cone neck diameter driven side

22 centering surface 49 moving cone disc width driving side

23 noise damping device 50 fixed cone disc width drive side

24 (center) holes 51 fixed cone disc width driven side

25 horizontal holes 52 moving cone disc width driven side

26 Turned groove 53 Small sliding fit seat length drive side

27 teeth structure 54 large sliding fit seat length drive side

28 connecting holes 55 large sliding fit seat length driven side

29 empty chambers/cavities 56 small sliding fit seat length driven side

30 Control hole 57 Compression spring

58 Sheet metal part (sliding rail for seal) 66 Second sliding fit seat

59 sealing device 67 support

60 Centrifugal Oil Cover 68 Housing Basic Structure

61 Flange parts 69 Hydraulic pump

62 Sliding fit seat 70 Sealing device

63 seal 71 weld

64 Starting unit 72 Centrifugal oil chamber

65 The first sliding fit seat 73 Distribution transmission device

74 (additional) centrifugal oil chamber

Claims (12)

1. The cone-disc winding contact device transmission (1) has a drive-side and a driven-side cone pair, each of which has a fixed cone (4, 46) and a moving cone (5, 33). These cones The disks are each arranged on a drive-side and a driven-side shaft (3, 41) and can be connected via a winding contact device (2) for torque transmission, characterized in that at least one moving cone (5, A sliding fit seat (65) of 33) is arranged in the radial inner region of the moving cone and at least one sliding fitting seat (66) of at least one moving cone (5, 33) is arranged in the moving cone in the radially outer region of . 2. The conical-disc winding contact device transmission according to claim 1, characterized in that: the movable conical disc (5, 33) has two sliding fitting seats. 3. The transmission of the conical-disc winding contact device according to claim 1, characterized in that the movable conical disc (5, 33) has three sliding fit seats. 4. The conical-disc winding contact transmission according to claim 1, characterized in that, when using radially outer sliding seats (66), a centrifugal oil cover (60) is formed. 5. The conical-disc winding contact transmission according to one of the preceding claims, characterized in that the sliding fit seat structure (65, 66) is arranged on the drive-side and/or driven-side conical-disc pairs (44, 45) superior. 6. The conical-disc winding contact gear transmission according to one of the preceding claims, characterized in that the radially outer sliding fit seat (66) passes through a sealing device (70) arranged axially adjacent to it be sealed. 7. According to one of the preceding claims, the conical-disc winding contact device transmission is characterized in that: the support (67) of the movable conical disc (5, 33) is arranged in the radially outer sliding fit seat (66) Radially inside. 8. According to one of the preceding claims, the conical-disc winding contact gear transmission is characterized in that the radially outer sliding fit seat (66) is formed using a single part (60) which is connected to the moving Cone disc (5,33) is connected. 9. The conical-disk winding contact device transmission according to claim 8, characterized in that said connection is a welded connection (71). 10. Conical-disc winding contactor transmission according to claim 8 or 9, characterized in that the part (60) is a centrifugal oil cover. 11. The conical-disc winding contactor transmission according to one of the preceding claims, characterized in that two centrifugal oil chambers (72, 74) are provided. 12. Motor vehicle, characterized in that it has a transmission according to one of the preceding claims.

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