Cross-Reference to Related Applications

This patent application is related to Italian Patent Application No. 102021000014756 filed on June 7, 2021, the entire disclosure of which is incorporated herein by reference .

Technical Field of the Invention

The present invention relates to a tensioner for a belt drive of a motor vehicle, in particular for a timing system.

State of the Art

Tensioners are known to comprise a fixed part provided with a pivot, a movable arm mounted on the pivot in a rotatable manner, and a pulley carried by the movable arm and intended to cooperate with the belt. In the case of a tensioner for a timing belt, the moving arm normally comprises an eccentric, pivot-mounted element which houses a spring. There is also a damping element interposed between the fixed part and the moving arm, typically defined by a bushing, so as to dampen the oscillations of the moving arm.

The "installation position" is defined as the maximum deviation of the arm from the belt, which can therefore be easily installed in this situation, and the "nominal position" as the static equilibrium of the tensioner under the action of the elastic force of the spring and the belt in nominal conditions.

Typically, tensioners damp vibration symmetrically, i.e. they exhibit a substantially equal hysteresis cycle during the loading and unloading phases of the tensioner. A limitation of such tensioners is the necessary presence of a certain angular distance between the nominal position and the installation position, in order to permit the belt installation. However, in some timing belt drives it is necessary, during operating conditions, to limit the angular displacement of the arm towards the installation position so as to keep it below a certain angle and prevent belt tooth jump. This may be achieved, in known tensioners, via a one way clutch interposed between the arm and the fixed part of the tensioner. This brings about additional components.

The purpose of the present invention is to realize a tensioner that can overcome the above problems.

Subject and Summary of the Invention

The above purpose is achieved by a tensioner as claimed in claim 1.

The tensioner according to the invention features a dual damping system, wherein the second non-symmetrical damping system provides a virtual stop to the arm after the nominal position in direction to the installation position. The tensioner allows the manual rotation of the arm until the installation position. However, during the engine operating conditions, it permits only a predetermined amount of arm angular displacement in direction opposite to the belt tensioning direction. Brief Description of the Drawings

For a better understanding of the present invention, a preferred form of implementation is described below, by way of non-limiting example and with reference to the accompanying drawings, wherein: Figure 1 is an axial sectional view of a tensioner according to the present invention;

Figure 2 is an exploded perspective view of the tensioner of Figure 1;

Figures 3, 4, 5 are partial sectional views of the tensioner of Figure 1 in various operating positions; and

Figure 6 is a graph representing the torque of the tensioner as a function of the angle of rotation of a tensioner arm.

Detailed Description of Preferred Embodiments of the

Invention

With reference to Figure 1, there is indicated by 1 a tensioner made according to the present invention and suitable for use in a timing system of an engine. The tensioner 1 includes a fixed part 2 provided with a base 3 and a pivot 4 fixed with respect to the base 3 and defining an axis A, an eccentric arm 5 rotatable about the pivot 4, and a pulley 6 rotatable about an axis B defined by the arm 5 and configured to cooperate with a timing belt (not shown).

The tensioner 1 further includes a first spring-damper assembly 7 and a second spring-damper assembly 11 interposed between the fixed part 2 and the arm 5, as will be further described below. In the present description, a "spring- damper assembly" is defined as an assembly comprising an elastic element, such as a spring, and a damping element, such as a bushing, acting jointly.

The base 3 is essentially an annular plate of axis A. The base 3 has, at an inner annular edge thereof, a first annular projection 12, and at an outer annular edge thereof, a second annular projection 13, extending in an axial direction. The pivot 4 is planted inside the first annular projection 12 and defines with the base 3 a fixed part 2 apt to be fixed to the engine by means of a screw of axis A (not illustrated) mounted through the pivot 4, and a projection

14 extending from the base 3 in an axial direction opposite to the projections 12, 13 and apt to engage with a seat obtained in the engine. The first spring-damper assembly 7 includes a primary spring 15 interposed between the fixed part 2 and the arm 5, and a damping bushing 16.

The arm 5 includes an eccentric portion 17 rotatably mounted on an end portion of the pivot 4 opposite the base 3 via the damping bushing 16 and having an inner cylindrical

A-axis surface in radial contact with the damping bushing 16 and an outer cylindrical B-axis lateral surface on which the pulley 6 is mounted by means of a bearing 21.

The arm 5 further includes a bell-shaped portion open to and facing the base 3, which is connected to the eccentric portion 17 by a radially annular intermediate wall 22. The bell-shaped portion bounds a chamber 23 with the base 3.

The primary spring 15 is a helical torsion spring, for example of circular cross-section, housed in the chamber 23 coaxially with the pivot 4 and axially interposed between the base 3 and the intermediate wall 22 of the arm 5. The primary spring 15 includes a first end 24 and a second end 25 directed radially. The first end 24 of the primary spring

15 engages in a radial through-seat 18 of the second annular projection 13 of the base 3. The second end 25 of the primary spring 15 engages a radial through-hole (not shown) of the bell-shaped portion of the arm 5.

The second spring-damper assembly 11 is housed in the chamber 23 axially between the base 3 and the intermediate wall 22 of the arm 5, and radially between the pivot 4 and the primary spring 15. The second spring-damper assembly 11 includes a fixed casing 26, a secondary spring 27, and a bushing 31. The casing 26, forming part of the fixed part 2, is cup shaped and includes an annular base portion 28 fitted on the pivot 4 and rigidly coupled to the base 3, for example by an axial projection 29 of the base 3 engaging a corresponding seat of the base portion of the casing 26, and an outer A- axis tubular wall 30 extending axially from the base portion toward the intermediate wall 22 of the arm 5. The casing 26 can be made of a variety of materials, for example, plastic material (possibly with an internal metal insert to increase mechanical strength), metallic material, or composite material.

The bushing 31 is a conical damping bushing mounted on the pivot 4 and comprising a conical wall 39 and an annular end flange 40 extending radially outward from the conical wall 39. The conical wall 39 has a radially inward surface 32 and a radially outward surface 33. The surface 32 is in radial contact with an intermediate portion of the pivot 4, and has a series of axial grooves 41. The surface 33 of the bushing 31 is substantially conical with a decreasing diameter from the flange 40 toward the base portion 28 of the casing 26, and adapted to interact with the secondary spring 27. The bushing 31 is axially cut along one of its generators so as to be radially flexible. On the side opposite the cut with respect to the axis A, the bushing 31 has an axial projection 38, preferably provided with an axial groove 42, which extends in the opposite direction with respect to the base 3 and is housed with angular clearance within a circumferential groove 44 of the arm 5.

The secondary spring 27 is a helical torsion spring, for example of rectangular cross-section, coaxial to the pivot 4 and mounted around the bushing 31, being radially interposed between the conical wall 39 of the bushing 31 and the tubular wall 30 of the casing 26. The secondary spring 27 includes a first end 34 and a second end 35 directed axially. The first end 34 of the secondary spring 27 engages an axially directed through-hole of the base portion 28 of the casing 26. The second end 35 of the secondary spring 27 engages an axial seat of the bushing 31, preferably in the axial groove 42 of the projection 38. The secondary spring 27 is mounted within the casing 26 with a radial preload.

The primary spring 15 exerts on the arm 5 such an elastic torque as to maintain the arm 5, in the absence of loads on the pulley 6, in a "free arm" position defined by the contact of the projection 38 of the bushing 31 with a first end 45 of the groove 44 of the arm 5 itself.

The tensioner 1 (Figure 2) further includes a lock-up pin 37 for locking the arm 5 in an installation position rotated with respect to the "free arm" position against the spring action of the primary spring 15 and the secondary spring 27, as will be better described below. The lock-up pin 37 is housed in a radial through-hole 43 of arm 5 and holds arm 5 in a fixed relative position with respect to base 3 in an installation configuration of the tensioner 1. Finally, the tensioner 1 includes a dust-cover disc 36 interposed axially between an end of the arm 5 and a head of the pivot 4, and extending radially so as to protect the bearing 21.

The operation of the tensioner 1 is described starting from the "free arm" position defined, as mentioned above, by the contact of the projection 38 of the bushing 31 with the first end 45 of the groove 43 (Figure 3).

As the belt load increases, the arm 5 rotates, against the reaction torque of the primary spring 15, toward the installation position. The connection with angular clearance between the arm 5 and the bushing 31 is such that in a first part of the rotational movement of the arm 5 (section a in the graph of Figure 6), only the first spring-damping assembly 7 intervenes. In particular, the primary spring 15 acts, whose elastic deformation increases with the increase of the relative angle according to a linear relation, with a first slope defined by the stiffness of the primary spring 15.

Upon reaching a predetermined angle, dependent on the angular width of the groove 44, the arm 5 comes to hit tangentially against the projection 38 of the bushing 31 with a second end 46 of the groove 44 (Figure 4). The stiffness of the primary spring 15 may be chosen such that this condition coincides with a nominal position of the tensioner 1, determined by balancing the actions exerted by the primary spring 15 and the belt in nominal conditions.

From this position, as the belt load increases, the second spring-damper assembly 11 also acts. In particular, the secondary spring 27, of greater stiffness than the primary spring 15, acts too. Since the secondary spring 27 is preloaded and it is necessary to overcome this preload, the graph (Figure 6) presents a second section b with an initially very high slope that progressively decreases with the increase in the number of coils which, while contracting, detach themselves from the tubular wall 30, and therefore with the increase in the free length of deformation of the secondary spring 27, until a condition of complete detachment of the secondary spring 27 from the tubular wall 30. At a further increase in the relative angle, the bushing

31 rotates against the action of both springs 15, 27. The graph (Figure 6) therefore presents a third section c with a slope defined by the sum of the stiffnesses of the two springs 15, 27 acting jointly in parallel. During its own contraction, the secondary spring 27 progressively and smoothly wraps around the bushing 31, creating radial contact one coil after the other. Due to its axial cut, the bushing 31 in turn progressively tightens on the pivot 4. Therefore, a proportional radial damping is realized between the bushing 31 and the pivot 4.

Having described the behaviour of the tensioner 1 as a function of the load exerted by the belt, it is now appropriate to describe its actual behaviour in use.

In particular, the installation of the tensioner 1 on the engine takes place in the installation position (Figure 5), held stable by the lock-up pin 37.

After installation of the belt, the lock-up pin 37 is pulled out and the arm 5, under the action of the springs 15, 27, moves to its nominal position (Figure 4).

Under dynamic conditions, during engine operation, the two spring-damper assemblies 7, 11 realize an asymmetric damping of the motion of the arm 5 depending on the direction of such motion from the nominal position. In particular, when the arm 5 rotates towards the "free arm" position, only the first spring-damper assembly 7 intervenes and therefore the damping is determined solely by the damping bushing 16. Instead, when the arm 5 rotates towards the installation position, the second spring-damper assembly 11 also intervenes, thus resulting in an increase in both the elastic stiffness and the damping acting on the arm 5.

The asymmetric damping is evident with reference to Figure 6, wherein the curve shows different hysteresis levels resulting from the different dampings described above.

Upon examination of the characteristics of the tensioner 1, the advantages of the present invention are clear.

In particular, the tensioner 1 dampens the vibrations of the arm 5 asymmetrically. This is possible thanks to the second spring-damper assembly 11, which creates a region with high stiffness and damping due to the combination of the preloaded secondary spring 27 and the bushing 31.

Furthermore, the preload of the secondary spring 27 is controlled by the casing 26. For the same secondary spring 27, the smaller the diameter of the tubular wall 30 of the casing 26, the smaller the outer diameter of the secondary spring 27 and thus the greater its preload. Therefore, the preload of the secondary spring 27 can be easily controlled by a convenient choice of the diameter of the casing 26 or of the number of coils of the secondary spring 27.

This allows to use (and thus make) fewer components with respect to a tensioner having a one-way clutch.

In addition, during movement of the tensioner 1 from the nominal to the installation position, the tensioner 1 absorbs most of the energy from the belt due to the resistant torque created by the secondary spring 27. In this way, less energy will remain to be damped by the friction of the bushing 31 into the pivot 4, reducing the wear of the bushing 31 and improving its life. This is advantageous with respect to known one-way clutch systems that dampen the movement by friction and not by a spring reaction torque, since the spring in a one-way clutch is not pre-loaded.

Furthermore, another advantage is that the secondary spring 27 with higher stiffness cooperates with the primary spring 15 when the tensioner 1 is in the installation position or any position in the range between the installation position and the nominal position and the belt load stops acting. This allows to increase the movement acceleration so as to make the tensioner 1 return to the nominal position faster when compared with a one-way clutch system. In this condition, the one-way clutch would stop acting, since it would only release the contact between the pivot and bushing and only the primary spring would drive the tensioner from the installation position to the nominal position.

Finally, it is clear that modifications and variations can be made to the tensioner 1 without going beyond the scope of protection defined by the claims.

For example, the primary spring 15 may not be circular in cross-section and/or the secondary spring 27 may not be rectangular in cross-section and be e.g. circular.

Further, the secondary spring 27 may be of such high stiffness that it achieves its purpose without needing to be preloaded or may have a different number of coils.