IT201600077370A1 - FLYWHEEL DOUBLE MASS AND SPIRAL SPRING PERFECTED - Google Patents

Description

"FLYWHEEL WITH DOUBLE-MASS AND SPIRAL SPIRAL PERFECTED"

The present invention relates to a dual mass flywheel and more particularly to an improved dual mass flywheel and spiral spring.

It is known to use double mass flywheels comprising a primary mass driven in rotation by a crankshaft of an internal combustion engine, a secondary mass capable of being connected to an input shaft of a gearbox and a filtering unit for connect the primary mass to the secondary mass in an elastic way.

It is known that the filtering unit comprises at least two elastic units arranged in series and having different stiffnesses. 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 be constituted respectively by a plurality of spiral springs interposed between the primary mass and the intermediate mass and by a plurality of circumferential helical springs interposed between the intermediate mass and the secondary mass.

In some operating conditions of the engine, such as for example at idle and particularly when the accessories (air conditioning compressor and generator) are activated, unwanted noise can occur due to torsional vibrations transmitted from the crankshaft to the dual mass flywheel and consequently to the transmission.

To reduce such torsional vibrations, it is also known that at least one of the elastic units comprises damping devices.

The object of the present invention is to provide an improved dual mass flywheel in which the aforementioned damping devices are made in a simple, economical and compact way.

The above object is achieved by a flywheel according to claim 1.

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 attached drawings in which:

Figure 1 is a functional diagram of a dual mass flywheel according to the invention;

Figure 2 is an exploded perspective view of the flywheel object of the invention with parts removed for clarity; Figure 3 is a first axial view, partially sectioned, of the flywheel of Figure 2;

Figure 4 is a second partially sectioned axial view of the flywheel of Figure 2;

figure 5 is a section along the line V-V of the flywheel of figure 4;

figure 6 is an enlarged view of a detail of figure 5; And

figure 7 is an axial view, partially sectioned along the line VI-VI of the flywheel of figure 5.

Figures 2 and 4 show a dual mass flywheel 1 of axis A and comprising a primary mass M1 capable of being connected to a drive shaft of an internal combustion engine (not shown), a secondary mass M2 capable of being connected to a shaft input of a transmission (not shown), for example by means of a clutch, and a filtering unit 3 which connects the masses d M2 together. The flywheel 1 has an axis A coinciding with the axis of the driving shaft and the transmission shaft.

The filtering unit 3 is axially interposed between the primary mass M1 and the secondary mass M2e essentially comprises a first intermediate mass M3, a second intermediate mass M4 and elastic and damping means which connect the two intermediate masses M3, M4 to each other and respectively with the masses and M2 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 integrally connected to the hub 10. An annular toothed crown 9 extending parallel to the axis A is fixed to the radially outer edge of the disc 11. cantilevered towards the secondary mass M2e internally delimited by an internal annular surface 9a.

The primary mass is connected to the first intermediate mass M3 by means of a pair of metal spiral springs 4 with quadrangular section, having respective internal ends 4a fixed in diametrically opposite positions to the first intermediate mass M3 and respective external ends 4b constrained to respective fixed spring-holder elements 14 in positions diametrically opposite to the surface 9a and therefore integral with the disc 11. These spring-carrying elements 14 have an arcuate shape and each have a slot 14a in which the end 4b of a respective spring 4 is integrally engaged. equivalent stiffness Kss have the same sense of winding and have a number of turns between 0.5 and 2.

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

With reference to Figure 5, between the spiral spring 4 and the primary mass M1 there is also axially interposed, integral with the primary mass Ml a friction damper, preferably a friction disc 7. The latter is placed in axial contact with the coils of the spiral springs 4 in order to provide damping Tsstra the primary mass Mi and the first intermediate mass M3, due to the mutual sliding between the two.

The first intermediate mass M3 essentially comprises a spring-holder element 12 having a substantially cylindrical shape provided with substantially tangential seats (not shown) for the interlocking of the internal ends 4a of the spiral springs 4. Said spring-holder element 12 is supported radially in a rotationally free way rolling bearings 61 on primary mass Μ ,. Between the primary mass M1 and the first intermediate mass M3 one or more safety stops are conveniently provided to limit the maximum torsional deformation of the spiral spring 4. Conveniently (figures 5, 7 and 8) these safety stops consist of radial projections of the hub 10 and by respective internal teeth of the spring-holder element 12.

A tubular portion 12a extends towards the secondary mass M2 from the spring-carrying element 12. On the tubular portion 12a there is a plurality of external radial projections 16 each defining a circular seat 13 with a radial axis; on the bottom of each seat there is a hole 15 also with a radial axis and having a smaller diameter than the respective seat 13. Each seat 13 houses a sliding block 19 which is radially movable and loaded towards the outside by a cylindrical spring 20 housed in the respective hole.

An annular axial projection 12b extends from the tubular portion 12b, in recess with respect to the tubular portion 12a, on which a damper 25 is mounted. an open metal ring 31 mounted with radial forcing on the bushing 30 and rotationally coupled thereto by means of a pair of radial projections (not shown) which engage in corresponding holes in the ring 31. The ring 31 comprises at its ends respective radial turn-ups 33 external elements adapted to cooperate with stop elements 34 integral with the second intermediate mass M4 as described below.

The second intermediate mass M4 comprises an annular actuator disk 40 on whose inner edge 41 there are formed equally spaced recesses 22 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 a circular profile of axis A and respective side portions 22d adjacent to the sides 22a in relief with respect to the intermediate portion 22c and connected 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, respectively minor and major, according to whether the shoes 19 cooperate with the intermediate portion 22c or with the lateral portion 22d.

Assuming that a shoe 19 is disposed equally angularly with respect to the sides 22a, the angle between each shoe 19 and each of the sides 22a will be equal to <a> / 2 'where a represents the total angular travel of the shoes 19 within the openings 22 . This stroke has a value of less than 30 °, preferably less than 24 °, and conveniently of 16 ° and is formed by a central section σ of for example 8 ° defined by the run of the shoe 19 against the central portion 22c and by respective sections end ε equal for example to 4 ° defined by the run of the shoe 19 against each of the side portions 22d.

The disc 40 further comprises a plurality of spokes 42 extending radially from the disc 40 itself and arranged equally angularly spaced, preferably four by two by two diametrically opposite. The disk 40 is free to rotate within a volume 43 enclosed by two shells 44, 45 integral with the secondary mass M2.

Friction means, preferably friction rings 47, are axially interposed between the disc 40 and the shells 44, 45, able to generate a damping between the disc 40 and the secondary mass M2 thanks to the sliding of the disc 40 within the friction rings 47 during its rotation around the axis A.

In particular, with reference to Figure 6, a first friction ring 47a is interposed between the disc 40 and the secondary mass M2 and a second friction ring 47b is interposed between the disc 40 and a collar 80. The disc 40 is therefore axially comprised between the friction rings 47a, 47b.

The collar 80 is preferably annular and comprises a horizontal portion 81, in contact with the second friction ring 47b and a vertical portion 82 extending from an internal radial end of the horizontal portion 81 towards the spiral springs 4 and connected to it.

The collar 80 is circumferentially integral with the shell 44 by means of the coupling between a projection 83 having the shape of an arc extending axially from a substantially central portion of the horizontal section 81 in a seat 84, of a shape substantially complementary to the projection 83, obtained on the shell 44 The coupling of the projection 83 in the seat 84 prevents relative rotations between the collar 80 and the shell 44 but allows a relative axial movement between the two.

The shell 44 further comprises a plurality of elastic means configured to provide an axial load on the collar 80 (see Figures 2, 5, 6 and 7).

Preferably, these elastic means are foil bending springs 70 arranged circumferentially around the axis A, even more preferably they are four foil bending springs 70 arranged equally spaced crosswise.

Preferably, each leaf bending spring 70 is formed integrally with the shell 44 and is preferably obtained by blanking therefrom.

Each foil bending spring 70 essentially comprises at least one foil 73 projecting from the shell 44 in a radial direction towards the axis A and acting on the collar 80. Preferably each foil bending spring 70 comprises two of said foils 73.

The interlocking portion of the laminae 73 to the shell 44 is preferably joined by means of a connecting portion 74 configured to decrease any stress concentrations.

The axial load provided by the foil bending springs 70 to the collar 80 is defined on the basis of the shape of the foils 73 and can be varied by changing the shape of the foils 73 or their number.

Furthermore, the laminae 73 preferably have a curvature 75 in the plane perpendicular to the axis A defining a concavity facing the disk 40. This curvature 75 allows to obtain a center of gravity of the laminae 73 placed above, that is towards the spiral springs 4, of the plane perpendicular to axis A. This displaced center of gravity allows, due to the centrifugal force generated at high rotation speeds of the flywheel 1, to obtain a slight displacement of the lamina 73 towards the collar 80. In this way, as it increases of speed the load applied by the lamina 73 to the collar 80 is constant if not slightly increased.

The axial load provided by the leaf bending springs 70 is uniformed by the collar 80 which distributes the force imparted by each leaf bending spring 70 on the friction ring 47b in a substantially uniform manner.

The pressure exerted by the collar 80 on the friction ring 47b allows the assembly consisting of the disc 40 and the friction rings 47 to be held in pack thanks to the reaction provided by the secondary mass M2 with respect to the force supplied by the leaf bending springs 70.

The leaf bending springs 70 are configured to provide the collar 80 with a load such as to generate a damping between the friction rings 47 and the disc 40 of less than 100 Nm, preferably between 40 and 60 Nm and advantageously of about 55 Nm.

The aforementioned stop elements 34 project axially from the disc 40 towards the damper 25. In the position of the damper 25 in which the flaps 33 are equidistant with respect to the respective stop elements 34, it is present between each of the flaps 33 and the respective stop element 34 is an angle β being the total angle of free rotation between the damper 25 and the actuator disk 40. The angle β is conveniently equal to the angle σ to allow the activation of the damper 25 when the shoes 19 interact with the lateral portions 22d in such a way that the respective dampings add up to form a damping. the leaf bending springs 70. The seats 51 are conveniently constituted by windows of the shells 44, 45 facing each other (Figure 2).

The elastic groups 50 comprise (Figure 6) a cylindrical helical spring 52 with a tangential axis and a pair of shoes 53 which house respective end portions of the spring 52 and are able to cooperate under the thrust of the spring 52 with respective ends 51a of the seats 51. The springs 51 have an equivalent stiffness much greater than the equivalent stiffness Kss of the spiral springs 4. The angular play δ between the shoes 53 of each elastic group 50, arranged in contact with the ends 51 of the respective seat 50, is conveniently equal at 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 in such a way that each of the spokes 42 is angularly equivocal with respect to the elastic groups 50, the angle between each of the spokes 42 and each of the elastic groups 50 will be equal to ^ l ^ i where γ represents the total angular play between the spokes 42 and the elastic groups 50.

The spokes 42 are able to cooperate with the elastic groups 50 causing the compression of the springs 52 and possibly the relative contact between the shoes 53 once the play δ between them has been recovered.

The secondary mass M2 comprises a disk 59 provided with a hub 58 supported in a rotationally free way on the hub 10 of the mass M1 thanks to rolling bearings 62. The shells 44,45 integrally constrained to the disk 59 are also part of the secondary mass M2. The disc 59 is adapted to be connected in a conventional way to a vehicle clutch.

With reference to Figure 1, the filtering unit 3 can be considered to consist of three stages in series with each other:

- A filtering stage F comprising:

- an elasticity Ksstra the primary mass M1 and the first intermediate mass M3 due to the overall stiffness of the spiral springs 4 and

- a damping Tss generated by the sliding of the spiral springs 4 on the friction disc 7; - A minimum stage comprising:

- a damping TG1 acting between the intermediate masses M3 and M4, due to the interaction between the pad 19 and the surfaces 22b, 22d as well as to the effect of the damper 25 and

- a clearance Gl r defined by the angle a between the runners 19 and the openings 22

A starting stage comprising:

- a damping T3a between the second intermediate mass M4 and the secondary mass M2 due to the interaction between the shells 44, 45 and the friction means 47, which are kept in contact with each other by means of the load provided by the leaf bending springs 70,

- a clearance G3, defined by the angle γ,

- a K4 elasticity in series with the G3 clearance, due to the springs 42 and

- a damping 7 in series with the clearance G3 and in parallel with the elasticity K1 due to the friction of the movement of the shoes 53 in the respective seats 51 during the compression of the springs 52. exchange.

The operation of flywheel 1 is as follows.

In conditions of engine speed (stationary motion or variable speed but without sudden accelerations or decelerations) the input torque acting on the primary mass is transmitted by the spiral springs 4 (stiffness Kss) to the first intermediate mass M3.

The angular play Gl is recovered due to the presence of the torque, of a value greater than 70Nm, for example, thus making the damping TGl inoperative and the spokes 42 of the actuator disc 40 come into contact with respective shoes 53 of the elastic groups 50. Since the stiffness is much greater than the stiffness Kss, the deformations of the springs 52 are much less than those of the spiral springs 4.

Therefore the torsional oscillations of the crankshaft are absorbed substantially by the filtering stage F.

In engine idling conditions (no torque but relatively high torsional oscillations), the filtering group F and the starting group S, equipped with respective Ksse Kl r stiffnesses, behave as rigid systems with respect to the idling group I, which therefore absorbs the oscillations thus making the spiral springs inactive 4.

For oscillations having an amplitude smaller than the angle o, 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 σ, the damper 25 is inactive since the flaps 32 do not interact with the stop elements 34.

For oscillations of greater amplitude than the angle o, the shoes 19 cooperate with the side portions 22d of the openings 22 thus generating a second damping level 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 o, parallel to the interaction of the shoes 19 with the side portions 22d, one of the two flaps 33 (depending on the direction of rotation) 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 second intermediate mass M3, M4, conveniently greater than 2 Nm, and preferably equal to 4Nm. This damping is added to the second damping level defining a total TG1 damping value of for example 6-10 Nm.

At start-up and in particular operating conditions of travel, such as for example gear change errors or abrupt reversals or variations in torque, the torsional oscillations can reach even higher values and in particular greater than the angle a. In such situations, the torsional oscillations are transmitted to the actuator disk 40 due to the contact between the shoes 19 and the sides 22b.

Once the clearance G3 between the spokes 42 of the actuator disk 40 and the elastic groups 50 has been recovered, the spokes 42 come into contact with respective shoes 53 of the elastic groups 50 and compress the springs 52, the purpose of which is to avoid rigid collisions between the actuator disk 40 and the secondary mass M2. The shoes 53 of each elastic group 50 come into contact only in the case of extreme stresses rarely reached under normal operating conditions of the engine.

When the flywheel rotates at very high speed, for example 4000-5000 rpm, the centrifugal force acting on the spiral springs 4 determines an anomalous deformation which could produce a spurious rotation of the first intermediate mass M3 and the possible intervention of the safety in the absence of torque or low torque, and resulting in noise.

The containment elements 17 of the spiral springs 4, by containing the radial expansion of the coils of the springs, prevent the aforementioned drawback from occurring.

The containment elements 17 also ensure that the radial expansion of the springs does not cause anomalous stresses in the areas of the spring adjacent to the outer end areas 4b for contact with the spring-holders 14.

During the work steps described above, the leaf bending springs 70 keep the disc 40 and the friction discs 47 in contact with each other, allowing the generation of a damping of a predetermined value during the operation of the flywheel 1.

As the rotation speed increases, the load imparted to the collar 80 by the leaf bending springs 70 remains constant or increases slightly, at most by about 5 Nm, thanks to the displaced center of gravity due to the curvature 75 which causes a displacement of the leaf 73 towards the disc 40.

From the foregoing, the advantages of a flywheel according to the invention are evident.

The use of elastic means made integrally integral with the shell 44 of the secondary mass M2 allows to provide the actuator 40 with a damping of a predetermined value with a minimum axial bulk.

The fact that these elastic means are sheet bending springs 70 obtained by means of a shearing operation of the shell 44 makes this solution economical and easy to produce.

The curvature 75 places the center of gravity of the leaf 73 of the leaf springs 70 so as to ensure constant damping as the speed of rotation increases, if not slightly increased.

Finally, it is clear that modifications and variations may be made to the flywheel 1 which, however, do not depart from the scope of protection defined by the claims.

For example, the elastic means could be other types of springs, not just leaf bending springs as described, provided that they are made integrally with the shell 44 and of reduced axial bulk.

The leaf bending springs 70 could have a shape or be arranged in a different way from as described, or be present in different numbers without thereby changing their function.

Furthermore, the leaf bending springs 70 could be made as a separate body and subsequently made integral with the shell 44.

Furthermore, the leaf bending springs 70 could be present on the shell 45 or on both in shells 44 45.

Claims (1)

CLAIMS 1.- Dual mass flywheel comprising a primary mass (Mi), a secondary mass (M2) and a filtering unit (3) interposed between said primary mass (Μ4) and said secondary mass (M2), the filtering (3) comprising a starting unit (S) comprising an actuator element (M4) coupled to said secondary mass (M2) through the series of an angular clearance (G3) and at least one spring, characterized in that said group d The starting (S) comprises elastic means (70) made integrally with said secondary mass (M2) and configured to provide an axial load on friction means interposed between said secondary mass and said actuator element to generate a damping (T3) placed in parallel to said series of said angular play (G3) and said spring. 2.- Flywheel according to Claim 1, that said friction means comprise friction discs (47) interposed between said secondary mass (M2) and said actuator (M4), said axial load acting to maintain said friction discs ( 47) in contact with said actuator element (M4). 3.- Flywheel according to one of Claims 1 or 2, characterized in that said elastic means generate a damping of less than 100 N. 4.- Flywheel according to Claim 3, characterized in that said elastic means generate a damping of between 40 and 60 Nm. 5. Flywheel according to one of Claims 1 to 4, characterized in that said elastic means are a leaf bending spring (70). 6.- Flywheel according to Claim 5, characterized in that said secondary mass (M2) comprises two shells (44, 45) within which said actuator element (40) is housed, said leaf bending springs (70) being obtained entirely on one of said shells (44, 45). 7. Flywheel according to one of claims 5 or 6, characterized in that each of said leaf bending springs (70) comprises a leaf (73) extending in a radial direction from said shell. 8. Flywheel according to one of Claims 5 to 7, characterized in that said leaf bending springs (70) are obtained by blanking from said at least one element (44) integral with said secondary mass (M2). 9.- Flywheel according to one of claims 5 to 8, characterized in that said lamina (73) has a curvature (75) in a plane perpendicular to an axis of rotation (A) of said flywheel, said curvature (75) placing the center of gravity of said lamina (73) so as to obtain the displacement of said lamina (73) towards said disk (40) as the rotation speed of said flywheel increases.