TECHNICAL FIELD

The present invention concerns a double mass flywheel and more specifically an improved double mass flywheel with spiral spring .

BACKGROUND ART

It is known that double mass flywheels are used comprising a primary mass rotated by a crank shaft of an internal combustion engine, a secondary mass adapted to be connected to an inlet shaft of a gearbox and an elastic assembly to elastically connect the primary mass to the secondary mass. It is known that the elastic assembly comprises at least two elastic units arranged in series and having different rigidities. The two elastic units are connected by means of a structure that has a non-negligible intermediate mass . For example the two elastic units can respectively consist of a plurality of spiral springs interposed between the primary mass and the intermediate mass and a plurality of circumferential helical springs.

In some operating conditions of the engine, for example when idling and particularly when the ancillaries (compressor of the conditioning system and generator) are activated, undesired noise can occur due to torsional vibrations transmitted from the drive shaft to the double mass flywheel and consequently to the transmission.

DISCLOSURE OF INVENTION

The object of the present invention is to produce an improved double mass flywheel which solves the problems connected with the known flywheels described above. The above object is achieved by a flywheel according to claim 1. BRIEF DESCRIPTION OF THE DRAWINGS

For a better understanding of the present invention, a preferred embodiment is described below, by way of non- limiting example and with reference to the accompanying drawings in which:

figure 1 is a functional diagram of a double mass flywheel according to the invention;

figure 2 is an exploded perspective view of the flywheel subject of the invention with parts removed for clarity;

figure 3 is a first partially sectional axial view of the flywheel of figure 2;

figure 4 is a second partially sectional axial view of the flywheel of figure 2;

figure 5 is a section according to the line IV- IV of the flywheel of figure 2;

figure 6 is a partially sectional axial view according to the line VI-VI in figure 4;

figure 7 is a schematic cross section according to line VII- VII in figure 5; and

figure 8 is a cross section analogous to that of figure 7 representing an alternative embodiment of the flywheel.

BEST MODE FOR CARRYING OUT THE INVENTION

Figures 2 and 4 show a double mass flywheel 1 with axis A and comprising a primary mass adapted to be connected to a drive shaft of an internal combustion engine (not illustrated) , a secondary mass adapted to be connected to an input shaft of a transmission (not illustrated) , for example a clutch, and a filtering unit 3 which reciprocally connects the masses * ^¾ s and - ^>; .. The flywheel 1 has an axi s A coinciding with the axis of the drive shaft and of the transmission shaft.

The filtering unit 3 is axially interposed between the primary mass i and the secondary mass ·¾ and essentially comprises a first intermediate mass ⁱ , a second intermediate mass and elastic and damping means which connect the two intermediate masses ^i¾ x< ^A * reciprocally and with the masses and - ^<¾ s respectively as described in detail below. The primary mass comprises a hub 10, adapted to be connected to the drive shaft (not shown) and a disc 11 connected integral with the hub 10. An annular crown 9 is fixed to the radially external edge of the disc 11, said crown extending parallel to the axis A in an overhanging manner towards the secondary mass and internally delimited by an inner annular surface 9a.

The primary mass ^i¾ i. is connected to the first intermediate mass ·* by means of a pair of quadrangular section metal spiral springs 4, having respective inner ends 4a fixed in positions diametrically opposite the first intermediate mass ^¾ and respective outer ends 4b secured to respective spring holder elements 14 fixed in positions diametrically opposite to the surface 9a and therefore integral with the disc 11. Said spring holder elements 14 have an arched shape and are each provided with a slit 14a in which the end 4b of a respective spring 4 integrally engages. The springs 4, having an equivalent rigidity , have the same winding direction and a number of coils between 0.5 and 2.

On the surface 9a of the crown 9 a plurality of containing elements 17 are fixed for the coils of the spiral springs 4 which, in addition to the deformations connected with transmission of the torque, are also subject to a radial expansion deformation due to a centrifugal effect which begins to be effective at speeds of approximately 1500-2000 rpm and becomes important at high rotation speeds of the flywheel 1, for example 4000-6000 rpm.

Figures 7 and 8 show two possible variations for the containing elements 17.

Figure 7 shows a first variation comprising four arch-shaped containing elements 17 having an outer profile 17a configured to couple with the surface 9a and an inner profile 17b configured so as to reproduce the deformed profile of the spiral spring 4 at a combination of design speed and torque, for example 4000 rpm and 220 Nm. The containing elements 17 in this case are in diametrically opposite pairs and fixed to the surface 9a in the angular spaces between the two spring holders 14, in each of which two containing elements are housed.

The variation of figure 8 has, in addition to the containing elements of figure 7, two further containing elements 17 positioned internally between the coils of the spiral spring 4 and fixed to the disc 11 for example by means of threaded connections or pins. In this case the stop 17 has two profiles 17c, 17d, outer and inner respectively, configured so as to reproduce the deformed profile of the coils with which they cooperate respectively, at a combination of design speed and torque, for example 4000 rpm and 220 Nm.

With reference to figure 5, a friction damper, preferably a friction disc 7, is furthermore axially interposed, integral with the primary mass ^iV - ^f i , between the spiral spring 4 and the primary mass ^iS' *s.. Said damper is positioned m axial contact with the coils of the spiral springs 4 in order to provide a damping - ^f ΐΐ between the primary mass i and the first intermediate mass , due to the reciprocal sliding between the two . The first intermediate mass ^{; i} s comprises essentially a spring holder element 12 having a substantially cylindrical shape provided with seats (not shown) substantially tangential for interlocking of the inner ends 4a of the spiral springs 4. Said spring holder element 12 is radially supported in a rotationally free manner by rolling bearings 61 on the primary mass - ^j i . Between the primary mass ^iS¾i i and the first intermediate mass one or more safety stops 63 are conveniently provided to limit the maximum torsional deformation of the spiral spring 4. Conveniently (figures 5, 7 and 8) said safety stops 63 consist of radial projections 64 of the hub 10 and respective inner teeth 65 of the spring holder element 12. From the spring holder element 12 a tubular portion 12a extends towards the secondary mass ¾ . On the tubular portion 12a a plurality of outer radial ridges 16 are provided each defining a circular seat 13 with radial axis; on the bottom of each seat a hole 15 is provided, also with radial axis and having a smaller diameter than the respective seat 13. Each seat 13 houses a radially movable shoe 19 loaded towards the outside by a cylindrical spring 20 housed in the respective hole . From the tubular portion 12a an annular axial ridge 12b extends, recessed with respect to the tubular portion 12a, on which a damper 25 is mounted. The damper 25 comprises substantially a C-shaped bushing 30 cooperating by sliding with the outer surface of the axial ridge 12b and an open metal ring 31 fitted by radial forcing on the bushing 30 and rotationally coupled to the same by means of a pair of radial ridges (not shown) which engage in corresponding holes in the ring 31. The ring 31 comprises, at its ends, respective outer radial ridges 33 adapted to cooperate with stop elements 34 integral with the second intermediate mass -^* as described below . The second intermediate mass -^* comprises an annular actuator disc 40 on the inner edge 41 of which evenly spaced recesses 22 are obtained in which the respective shoes 19 of the first intermediate mass are free to move circumferentially . The recesses 22 are delimited by oblique sides 22a diverging towards the inside of the actuator disc 40 and have a bottom surface 22b having an intermediate portion 22c with circular profile and axis A, and respective lateral portions 22d adjacent to the sides 22a in relief with respect to the intermediate portion 22c and joined to it.

The shoes 19 are pushed by the respective springs 20 against the bottom surface 22b of the respective recesses; each spring 20 exerts different forces, smaller and greater respectively, according to whether the shoes 19 cooperate with the intermediate portion 22c or with the lateral portion 22d.

Assuming that a shoe 19 is arranged angularly evenly spaced with respect to the sides 22a, the angle between each shoe 19 and each of the sides 22a will be equal to where ^w represents the total angular stroke of the shoes 19 within the openings 22. Said stroke has a value of less than 30°, preferably less than 24°, and conveniently 16° and is formed by a central section σ for example 8° defined by the stroke of the shoe 19 against the central portion 22c and by respective end sections ^ equal for example to 4° defined by the stroke of the shoe 19 against each of the lateral portions 22d. According to a variation, not shown, the openings 22 can be coated in a layer of polymer material.

The disc 40 further comprises a plurality of spokes 42 extending radially from the disc 40 and arranged angularly evenly spaced from one another, preferably four, in diametrically opposite pairs. The disc 40 is free to rotate within a volume 43 enclosed by two shells 44,45 integral with the secondary mass ^? '½ . Between the disc 40 or the shell 44 and the secondary mass ^m z friction rings 47 are axially interposed, adapted to define a damping between the second intermediate mass and the secondary mass.

The above-mentioned stop elements 34 extend axially in an overhanging manner from the disc 40 towards the damper 25. In the position of the damper 25 in which the ridges 33 are e^venly spaced from the respective stop elements 34, an angle 7

' 2 is present between each of the ridges 33 and the respective stop element 34, β being the total free rotation angle between the damper 25 and the actuator disc 40. The angle β is conveniently equal to the angle σ to allow activation of the damper 25 when the shoes 19 interact with the lateral portions 22d so that the respective dampings add up to form a damping TGI .

The shells 44,45 enclose, in seats 51 obtained therein, a plurality of elastic assemblies 50, preferably four, arranged circumferentially and angularly evenly spaced from one another. The seats 51 conveniently consist of windows of the shells 44,45 facing each other (figure 2) . The elastic assemblies 50 comprise (figure 6) a cylindrical helical spring 52 with tangential axis and a pair of shoes 53 which house respective end portions of the spring 52 and are adapted to cooperate under the thrust of the spring 52 with respective ends 51a of the seats 51. The springs 51 have an equivalent rigidity &t. much greater than the equivalent rigidity ¾Ϊ of the spiral springs 4. The angular clearance δ between the shoes 53 of each elastic assembly 50, arranged in contact with the ends 51 of the respective seat 50, is conveniently equal to 4°. The shoes 53 are also adapted to cooperate with the spokes 42 of the disc 40. Assuming that the actuator disc 40 is arranged so that each of the spokes 42 is angularly evenly spaced from the elastic assemblies 50, the angle between each of the spokes 42 and each of the elastic assemblies 50 will be equal to /2 , where represents the total angular clearance between the spokes

42 and the elastic assemblies 50.

The spokes 42 are adapted to cooperate with the elastic assemblies 50 determining the compression of the springs 52 and if necessary the relative contact between the shoes 53 once the clearance δ between them has been taken up.

The secondary mass ^;¾ s comprises a disc 59 provided with a hub 58 supported in a rotationally free manner on the hub 10 of the mass by the rolling bearings 62. Also the shells 44,45 integrally secured to the disc 59 form part of the secondary mass ^m 2~ The disc 59 is adapted to be connected m a conventional manner to a clutch of the vehicle. With reference to figure 1, the filtering unit 3 can be considered to be composed of three stages in series:

a filtering stage F comprising:

^■ an elasticity ^& S_? between the primary mass and the first intermediate mass ^i1¾' s due to the overall rigidity of the spiral springs 4 and

^■ a damping ⁱ SS generated by sliding of the spiral springs 4 on the friction disc 7 ;

an idling stage comprising:

^■ a damping * ¾ acting between the intermediate masses ^i¾ s and ·¾ , due to the interaction between the shoe 19 and the surfaces 22b, 22d and also to the effect of the damper 25 and a clearance defined by the angle a between the shoes 19 and the openings 22 a starter stage comprising: ^■ a damping s between the second intermediate mass -^* and the secondary mass - ^:Vi s due to the interaction between the shells 44,45 and the friction rings 47, and

" a clearance G3, defined by the angle γ,

^■ an elasticity in series with the clearance G3, due to the springs 42 and

^■ a damping *s in series with the clearance G3 and parallel to the elasticity due to the friction of the movement of the shoes 53 in the respective seats 51 during compression of the springs 52.

Two rigidities *½ , ¾ ^■ representing the equivalent rigidities of the drive shaft and the gearbox respectively.

Operation of the flywheel 1 is as follows.

With the engine in a steady state (fixed or variable speed but without sudden accelerations or decelerations) the inlet torque acting on the primary mass ⁱⁱ i is transmitted from the spiral springs 4 (rigidity ^A is ) to the first intermediate mass

The angular clearance Gl is taken up by the existing torque, with value higher, for example, than 70Nm, thus making the damping ^■ * ¾ ineffective and the spokes 42 of the actuator disc 40 come into contact with respective shoes 53 of the elastic assemblies 50. Since the rigidity ¾ is much greater than the rigidity *¾ ^■ , the deformations of the springs 52 are much less than those of the spiral springs 4.

Therefore the torsional oscillations of the drive shaft are substantially absorbed by the filtering stage F. With the engine in idling conditions (torque absent but relatively high torsional oscillations), the filtering assembly F and the starter assembly S, provided with respective rigidities -¾r e ¾ , behave like rigid systems with respect to the idling assembly I, which therefore absorbs the oscillations thus making the spiral springs inactive 4.

For oscillations with amplitude smaller than the angle σ, the shoes 19 cooperate with the central portions 22c of the openings 22 thus generating a first damping level conveniently lower than 3 Nm, and preferably equal to 1-2 Nm. In the example illustrated in which the angle β is equal to the angle a, the damper 25 is inactive since the ridges 32 do not interact with the stop elements 34.

For oscillations with amplitude greater than the angle σ, the shoes 19 cooperate with the lateral portions 22d of the openings 22 thus generating a second level of damping conveniently greater than 2 Nm, and preferably equal to 4-7 Nm. In the example illustrated in which the angle β is equal to the angle σ, parallel to the interaction of the shoes 19 with the lateral portions 22d, one of the two ridges 33 (according to the rotation direction) cooperates with the respective stop element 34 and tends to close the ring 31 on the bushing 30 increasing the forcing thereof on the surface 12b. An additional damping is therefore generated between the first and the second intermediate mass , , conveniently greater than 2 Nm, and preferably equal to 4Nm. Said damping is added to the second damping level defining a total damping value equal for example to 6-10 Nm. At starting and in particular operating conditions, for example gear change errors or sudden torque inversions or variations, the torsional oscillations can reach even higher values and in particular greater than the angle a. In these situations the torsional oscillations are transmitted to the actuator disc 40 due to the contact between the shoes 19 and the sides 22b. Once the clearance G3 between the spokes 42 of the actuator disc 40 and the elastic assemblies 50 has been taken up, the spokes 42 come into contact with respective shoes 53 of the elastic assemblies 50 and compress the springs 52, the purpose of which is to avoid rigid impact between the actuator disc 40 and the secondary mass M2. The shoes 53 of each elastic assembly 50 come into contact only in the event of extreme stress, which rarely occurs in the normal operating conditions of the engine.

When the flywheel rotates at a very high speed, for example 4000-5000 r.p.m., the centrifugal force acting on the spiral springs 4 causes an abnormal deformation thereof which could produce a spurious rotation of the first intermediate mass ⁱ *s and possible intervention of the safety stops 63 in the absence of torque or at low torque, and with consequent noise.

The containing elements 17 of the spiral springs 4, containing the radial expansion of the spring coils, prevent occurrence of the above-mentioned drawback.

The containing elements 17 also ensure that the radial expansion of the springs does not cause abnormal stress in the areas of the spring adjacent to the outer end areas 4b due to contact with the spring holder 14.

From the above, the advantages of a flywheel according to the invention are obvious.

The introduction of a clearance ^tr i and a damping in the coupling between the two intermediate masses ^iV *¾« ^iSS 4 reduces the torsional oscillations transmitted to the gearbox when the engine is idling, bypassing the spiral springs 4.

In the preferred embodiment illustrated, the damping ^{i U} i is variable with the relative rotation angle between the intermediate masses . The increase in the damping as said angle increases allows management of the impact between the two intermediate masses *¾.■ ^■ ·¾ and therefore a consequent reduction in noise.

The use of the shoes in parallel sliding against the surfaces 22c, 22d and of the damper 25 allows high damping to be obtained in a compact inexpensive manner.

Lastly, it is clear that modifications and variations can be made to the flywheel 1 which do not depart from the protective scope defined by the claims.

For example, the angular clearance α, β, γ, σ, δ, ^< ~ can be varied or the shape or number of the elastic assemblies 50, of the containing elements 17 or of the spring holder elements 14 without modifying the function thereof.