This invention relates to a traction drive type continuously variable transmission used in automobiles and various industrial machines.
Many inventions have heretofore been made concerning the continuously variable transmission (CVT). Recently, the half toroidal type (
The problem for traction drive devices is to efficiently transmit large torque with size compaction. To obtain large transmission torque with size compaction, this may be attained by giving a large contact force; however, excessively large contact stress shortens life. To reduce the contact stress, if the device is designed to increase the contact area, the spin component in the contact region increases, lowering the transmission efficiency.
It is the half toroidal type and metal belt type that have solved these problems to a certain extent, not completely, however, and each has its drawbacks. The half toroidal type swings and changes speed in a state in which a power roller is pressed against an input/output disk surface. A large contact pressure acts thereon, and if no oil film is formed, seizure is caused. Avoiding this needs to improve the surface roughness. Processing a large spherical surface with high accuracy inevitably results in high cost. Further, in the direction of rotation, convexes contact each other, so that a large spin component is produced depending on the contact position, thus lowering the transmission efficiency. Structurally, this type is axially elongated, and cannot be mounted on FF vehicles. On the other hand, the metal belt type is in the form of a stack of a number of elements for easy bending, adapted to be pressed against a V-pulley to transmit torque. Basically, the pulley and the belt are in metal-to-metal contact, so that wear is unavoidable.
An object of the invention is to provide a traction drive type continuously variable transmission which has solved the above problems and which is capable of efficient transmission of large torque with size compaction.
The invention provides a traction drive type continuously variable transmission of novel construction, which is a transmission in the form of a combination of a coreless transmission ring and a V-pulley, the coreless ring being held from the outer periphery by guide rollers.
That is, according to an embodiment of the invention, the traction drive type continuously variable transmission comprises a first input/output shaft rotatably supported by a casing, a second input/output shaft rotatably supported by the casing, a V-pulley consisting of a pair of pulley members supported by the first input/output shaft and having a V-groove whose width is variable, a ring engaged with the V-pulley and supported around its outer periphery, and a mechanism for moving the ring around the axis of the second input/output shaft. As compared with the half toroidal type, the traction drive type continuously variable transmission of the invention has the following advantages. Even if the contact area is increased, the efficiency will be high with little spin component. No spherical processing as in the half toroidal type is required. The transmission is shorter in axial length than in the half toroidal type and applicable to FF vehicles. Further, as compared with the metal belt type, it is simple in construction and has oil films formed in the interface to the pulley, and compared with the metal belt type, no wear is produced, so that it has a long life.
These and other objects and features of the invention will become more apparent from the following description with reference to the drawings.
a is an exploded perspective view illustrating a face clutch;
b is side view showing the operating procedures of the face clutch;
a is a sectional view showing the prior art;
b is a perspective view showing the prior art;
a is a longitudinal sectional view of the tilting pad bearing;
b is an enlarged view of an area b of
a is a conceptual view;
b is an enlarged sectional view;
Embodiments of the invention will now be described with reference to the accompanying drawings.
The first input/output shaft 6 has an input/output gear 2 fixed thereto. The input/output gear 2 is in mesh with the toothed ring 3. The sectional shape defined by the lateral surfaces of the toothed ring 3 substantially coincides with that of the groove of the V-pulley 4. The toothed ring 3 has teeth for meshing with the teeth of the input/output gear 2, and a smooth cylindrical guide surface 8, at which it is contacted with a guide roller 1. As guides for the toothed ring 3, since the load of contact with the toothed ring 3 is low, employment may be made, in addition to the guide roller 1 adapted to roll in contact with the outer peripheral surface of the toothed ring 3 as shown, of a slide bearing (shoe) adapted for slide contact with the toothed ring 3. As shown in
The second input/output shaft 7 has a spline shaft section 12, on which a pair of pulley members 4 are spline-fitted. The pulley members 4 are movable axially of the input/output shaft 7. Each pulley member 4 has a groove width adjusting mechanism 9. The groove width adjusting mechanism 9 includes a pair of face cams 21 and 22 supported coaxially with the input/output shaft, and a thrust bearing 15. Of the pair of face cams, the movable cam 21 is movable axially of the input/output shaft and contacted with the pulley member 4 through the thrust bearing. The fixed face cam 22 is fixed to the casing 10.
The pair of face cams are contacted with each other through slopes such that relative rotation provides movement toward or away from each other. Interposing balls between the inclined surfaces will smooth the movement. The face cams illustrated in
Gears 23 are fixed to the arm 13 and supported coaxially with the turning shaft of the arm 13. Further, the movable face cam 21 has teeth on the outer periphery, meshing with the gear 23 in a meshing section denoted by the reference character 20 in
The pair of gears 23 are integrated together by a connecting section 18. Therefore, the pair of gears 23 rotate only in synchronism. As a result, the movable face cams 21 in the right and left groove width adjusting mechanisms 9 in
Since the toothed ring 3 is restrained from the outer periphery by the three or more guide rollers 1, it can be rotated even without a central shaft, (coreless roller). The guide rollers 1 are connected by the arm 13, so that by turning the arm 13 the center of rotation of the toothed ring 3 can be moved around the center O1. Therefore, the teeth cut in the outer periphery of the toothed ring 3 are always in mesh with the teeth of the input/output gear 2. If the pulley members 4 and the arm 13 are controlled so that there is no clearance produced between the toothed ring 3 and the pulley members 4, movement of the toothed ring 3 around the center O1 changes the contact point between it and the pulley members 4; thus, the speed of the pulley members 4 can be continuously varied with respect to a given number of revolutions of the input/output gear 2. In this manner, the so-called CVT is constructed.
If the input/output shaft 7 supporting the pulley members 4 is used as an input shaft, the pressed-in state of the toothed ring 3 is a speed reduction state. If the transmission torque is the same, the clamping force due to the pulley members 4 with the toothed ring 3 pressed in should be large, and reversely it may be small if the contact point between the toothed ring 3 and the pulley members 4 is on the larger diameter side. When the bending stresses in the pulley members 4 due to the clamping force are considered, this method capable of reducing the clamping force, and this method in which the pulley members 4 is used for input is better than a method in which they are used for output.
When a rotational force is inputted from the pulley members 4 in the direction of arrow of
An embodiment shown in
A gear 30 is attached to the feed screw shaft 28. And, the feed screw shaft 28 is rotationally driven through the gear 30 by an unillustrated driving means. The right and left feed screw shafts 28 symmetrically operate to change the spaced distance (groove width) between the pair of pulley members 4, with the result that the speed change ratio is changed. That is, the rotation of the feed screw shafts 28 does not change the central positions of the pulley members 4 but changes the groove width alone. Thus, the feed screw shafts 28 correspond to the previously mentioned pulley width adjusting mechanisms 9 in that the feed screw shafts 28 change the groove width of the pair of pulley members 4 by their axial movement. In addition, common steel balls are interposed between the pulley members 4 and the shaft 7 to allow smooth axial relative movement, while a whirl-stop is provided in the direction of rotation.
As shown in
As for control methods for the force P, most simply, as
The contact region between the toothed ring 3 and the pulley members 4 assumes a high surface pressure. Therefore, if an edge load is produced, this will result in early peeling. As measures against this, it is preferable to give an auxiliary curvature having a radius of curvature r to the toothed ring 3, as shown in
Installing a speed change ratio setting sensor 64 and measuring the position of the end surface of the gear 30 enables the groove width of the pulley members 4 to be indirectly measured. The groove width of the pulley members 4 is equivalent to the speed change ratio. Accordingly, signals from the sensor 64 indicate the speed change ratios, the speed change being realized by rotating the gear 30 by an unillustrated outside motor. To control the drive power of this outside motor, it is necessary to know a target seed change ratio and an actual speed change position.
It is clear that the lubricating oil used for traction drive exhibits a high traction coefficient. If the contact surface pressure lowers, the traction coefficient also lowers. For this reason, it is necessary to maintain a minimum contact surface pressure. In the case where the angle θ of the pulley members 4 is constant, if the position of contact with the toothed ring 3 approaches the outer diameter of the pulley members 4, the peripheral length of the contact region increases and the contact surface pressure lowers. To avoid this, it is recommendable to change the angle of the pulley members 4 as shown in
Next, an embodiment shown in
The end surfaces of the pulley member 4 and feed screw shaft 28 face a common oil pressure chamber 46, which is sealed by a seal S. The lubricating oil used in the traction drive type continuously variable transmission is pressurized by a hydraulic pump and supplied to the right and left oil chambers 46. The pressure receiving area of the pulley member 4 is smaller than that of the feed screw shaft 28; therefore, this is less effective in reducing the load on the bearing is small but this is effective in reducing the thrust load between the feed screw shaft 28 and the nut 26. Further, this is also effective in downsizing the motor. To move the pulley members 4, the feed screws 28 are rotationally driven through the drive gears 30 by the gear device. At this time, the oil pressure supplied in such a manner that the drive power reduces is controlled. If an electric motor is used as a source for the drive power, the motor current is equivalent to torque and is easy to use for control.
The axial force acting on the pulley members 4 is determined by the transmission torque and speed change position of the traction drive type continuously variable transmission. At a high speed reduction position, the force with which the toothed ring 3 is pressed into between the pulley members 4 increase. The input torque is approximately proportional to the absolute pressure in the suction manifold of the engine. Accordingly, it is possible to calculate an optimum oil pressure by using the speed change position and suction pressure of the traction drive type continuously variable transmission obtained by the sensor 64.
a and 15b are conceptual views similar to
F3RG cos 20°=F1RT (1)
The force PSUM for pressing the toothed ring 3 into between the pulley members 4 is obtained from the pressing force from the guide rollers 1 and from Formula 1, where cos 20°≈1.
PSUM=P+F1RT/RG (2)
When the opening angle of the pulley members 4 is 2 θ, the contact force Q between the pulley members 4 and the toothed ring 3 due to PSUM is as given in Formula 3.
2Q sin θ=PSUM (3)
When the traction coefficient is μ, since 2Qμ=F1, Formula 4 is obtained from Formulae 1, 2, and 3.
F1=μ(P+F1RT/RG)/sin θ (4)
The traction coefficient changes with lubricating oil, temperature and slip factor. The relation between the slip factor and the traction coefficient is as shown in
The clamping force exerted by the pulley members 4 changes in proportion to the transmission torque, a fact which is optimum for reduction of friction loss; therefore, an ideal state is that Formula 4 holds when P=0. Accordingly, from Formula 4, the optimum pulley angle is when the relation of Formula 5 holds. However, in this case, μMAX is employed as μ.
sin θ=μ(RT/RG) (5)
If the θ is too smaller than that given by the relation of Formula 5, the contact surface pressure rises more than is necessary, leading to the lowering of life and the increase of torque loss. Reversely, if it is too large, slip increases in the traction section, and torque loss increases.
Generally, while the maximum traction coefficient of lubricating oil changes with the kind and temperature of lubricating oil, a practical maximum value μMAX us in the range of 0.1-0.07 and 1>RT/RG>0.8 or thereabout. Therefore, from Formula 5, the angle μ of the pulley members 4 providing a clamping force proportional to transmission torque is 7-3 degrees.
In the case where the stepping on the accelerator pedal is released or where the vehicle speed has increased, the drive force (F3) decreases. At this time, if the pulley member 4 are moved in the direction to narrow the groove width (shift up), the Q increases to press the toothed ring 3 outward to the outer periphery of the pulley members 4, with the speed reduction ratio shifting in the direction to decrease.
The toothed ring 3 is held by at least three guide rollers 1. It is necessary for the guide rollers 1 to keep the toothed ring 3 pressed against the pulley members 4 with a minimum force. The reason is that when the transmission torque is zero, the toothed ring 3 becomes unstable. As this pressing method, the use of spring force, oil pressure or electromagnetic force (motor or the like) is contemplated. In the case where the present traction drive type continuously variable transmission is applied to an automobile, it has been shown that self-preloading is possible during driving by properly designing the angle of the pulley members 4. During engine braking, however, the force acting from the output gear 2 to the toothed ring 3 acts in a direction away from the pulley members 4. Since this force changes mainly according to the speed change position, it is controlled by another control mechanism during application of engine braking.
During application of engine braking, the force acting from the output gear 2 to the toothed ring 2 acts in a direction away from the pulley members 4. The magnitude is 1/B of the maximum value during driving, B being 5-10. During application of engine braking, a pressing force twice said value has to be generated by oil pressure or another mechanism. That is, in Formula 4, if P≈2F1RT/BRG and RT≈RG, then P=2F1/B. The F1 changes according to the speed change position and increases on the high reduction-speed side.
When the speed change ratio is in the D range, this means that engine braking is not positively demanded. Accordingly, employed as the value of P at this time is F1 at a position where the speed change ratio corresponds to the top.
When positive engine braking is demanded, for example, when the engine is shifted to the third or second gear, P due to the speed change lever position will be added.
Consider the case where a maximum torque of 200 Nm is inputted into the present traction drive type continuously variable transmission. If RT=100 mm, then F1=2,000 N. If B=10 and RT≈RG, then P=2,000×0.2=400 N.
An embodiment shown in
In
Number | Date | Country | Kind |
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2003-32704 | Feb 2003 | JP | national |
2003-372635 | Oct 2003 | JP | national |
2003-410715 | Dec 2003 | JP | national |
This application is a National Stage entry of International Application No. PCT/JP2004/001156, filed Feb. 4, 2004, the entire specification claims and drawings of which are incorporated herewith by reference.
Filing Document | Filing Date | Country | Kind | 371c Date |
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PCT/JP04/01156 | 2/4/2004 | WO | 4/14/2006 |