DE102014006691A1 - Nonuniformity reduction through power split - characteristic design - Google Patents

Abstract

A torsional vibration damping arrangement for the powertrain of a motor vehicle, comprising an input to be driven for rotation about an axis of rotation (A) and an output portion, a coupling assembly connected to the output portion, the coupling assembly comprising a first input member, a second input member and an output member, and a torque transmission path for transmitting a total torque that passes between the input area and the output area, the torque transmission path from the input area to the coupling arrangement in a first torque transmission path and a parallel second torque transmission path is divided and a phase shifter assembly in the first torque transmission path comprising a vibration system with a a first spring arrangement about the axis of rotation (A) rotatable output element, wherein the first spring arrangement at least a first spring element, a second Spring element and a last spring element comprises, wherein the spring elements are connected in series, wherein the first spring element compared to the second spring element has a lower rigidity and that the second spring element compared to the last spring element has a lower rigidity.

Description

I. Figure Directory

1 spring work

1.1 Spring work - theoretical / real

1.2 Spring rework with frictionless spring

2 Friction conditions with spring plate / sliding block design - general

3 Forces on the shoe

4 Friction conditions with spring plate / sliding shoe version - Variant 1

5 Friction conditions with spring plate / sliding shoe version - Variant 2

6 Friction conditions with spring plate / sliding shoe version - variant 3

7 Characteristic design taking into account the spring arrangement

8th Characteristic - general

9 Characteristic - 5-stage

10 Characteristic - 4-stage

11 Characteristic - 2-stage

11.1 Characteristic - 2-stage - Variant 1

11.2 Characteristic - 2-stage - Variant 2

12 Characteristic-1-stage

13 Example of constructive implementation of a multi-level characteristic

13.1 constructive design of the power split

13.2 Representation of the spring control

13.3 Representation of the spring arrangement (outer damper and inner damper)

13.4 Representation of the single and total characteristic curves

14 Advantageous combination of number of planet gears and characteristic curve - 1

15 Advantageous combination of number of planet gears and characteristic curve - 2

16 Example characteristic: internal damping torque <Mmax

17 Example characteristic curve: internal damping torque ≧ Mmax

18 Example characteristic for thrust (internal damper torque ≦ Mmax_Schub

19 Example characteristic for train single-row

20 constructive example of external, axially adjacent spring sets integrated in a converter (for AT gearbox)

II. Item Number Directory

• deleted

III. Description, structure and function

1. Description

The requirements for reducing rotational nonuniformity in a vehicle powertrain are high and will continue to increase in future vehicle generations. One of the reasons for this is the legislature's increasing stricter CO2 emissions standards as well as rising customer demands.

The tightened fuel economy regulations will allow future engines to produce higher rotational irregularities. Reasons for this are z. B. higher mean pressures, downspeeding, downsizing, cylinder deactivation or any combination of measures.

For this reason and because of the customer demands for at least constant comfort, in particular with regard to the noise and vibration levels, performance-enhancing measures are required on all systems for reducing rotational nonuniformity.

State of the art

From the prior art torsional vibration damping arrangements for the drive train of a vehicle with two torque transmission paths, both of which emanate from the input area known (see DE 10 2012 218 729 A1 . DE 10 2012 214 361 A1 . DE 10 2011 007 118 A1 ). In this case, a torque transmission path via a phase shifter arrangement which extends a spring arrangement, for example with at least one bow spring and or also with at least one cylinder compression spring, whereas the other torque transmission path is rigid from the input area. By means of a coupling arrangement, the two torque transmission paths are brought together again and thus the split torque back together to form a torque, or destructively superimposed to superimpose by this overlay the torsional vibrations contained so that they are extinguished in the optimal case. This torsional vibration damping arrangement with two torque transmission paths, which can also be referred to as power split for a short time, can make a significant contribution to reducing torsional vibrations. The design of the power split or individual parameters such as the spring characteristic differs from the criteria of other decoupling systems (such as ZMS and ZMS with DAT). Basis of the invention are therefore design criteria for an advantageous design of the spring characteristic for the torsional vibration damping arrangement with power split, or for the phase shifter contained therein.

2nd construction

The spring characteristics of the power split can be constructed with bow springs, consisting of one or more outer springs, with one or more inner springs each with the same or different length and / or with cylinder compression spring, with one or more inner springs. For both types of spring, the spring sets (outer and inner spring) can be constructed with different or equal stiffnesses. Any combinations of both types of spring are equally possible. The use of sliding shoes and spring plates on all or individual springs (bow and cylinder springs) can be implemented in any combination.

• – ein Außendämpfer ohne Innendämpfer
• – ein Außendämpfer und radial innerhalb ein Innendämpfer
• – zwei Außendämpfer axial nebeneinander und ein Innendämpfer radial innerhalb der Außendämpfer,
• – ein Außendämpfer und radial innerhalb zwei Innendämpfer axial nebeneinander
• – zwei Innendämpfer axial nebeneinander
• – weitere Kombinationen aus radial und axial angeordneten Federsätzen ist ebenfalls vorstellbar

In addition, the required characteristic of one or more outer dampers z. B. axially next to each other 20 , from the combination outer damper with inner damper 13.1 - 13.4 or without external dampers with one or more internal dampers of any stiffness characteristics. The type of interconnection, series and / or parallel connection, individual spring groups depends on the required function and can also be combined arbitrarily constructive. The arrangement of the dampers can be implemented, for example, as follows:

• - an outer damper without inner damper
• - An outer damper and radially inside an inner damper
• Two outer dampers axially next to one another and an inner damper radially inside the outer dampers,
• - An outer damper and radially inside two inner damper axially adjacent
• - Two internal dampers axially next to each other
• - Other combinations of radially and axially arranged spring sets is also conceivable

Furthermore, when using cylinder compression springs, the order of arrangement of the different spring stiffnesses / masses (= sequence in the spring set, hard-medium-hard-soft-soft-hard), in the entire speed range, for the decoupling of importance.

All other measures to influence the friction of the spring set (lubrication, fliers, slideway plate, friction-reducing sliding inserts / coatings, rolling skids) can be used in addition to the measures described.

3rd function

3.1 Sequence influence of the springs in the spring set (here on the example of an outer damper with spring plates and sliding shoes)

In 1.1 is the spring stiffness for a frictionless (c1_theoretical) and a real, frictional (c1_real) spring as part z. B. an external damper shown. The area between the two springs represents the friction work of the spring, which arises between the spring and the Gleitbahnblech directly or between the spring and a sliding shoe. The friction / friction work that must be overcome leads to a real higher spring stiffness compared to a frictionless working spring.

Therefore, the available spring work of a real, frictional spring, at the same moment, is lower (compared to a frictionless spring). 1.2 ).

In 2 the installation situation of an outer damper with spring plates and sliding shoes is shown. The moment is introduced (static and alternating torque) on the primary flywheel (PSR) and discharged, to the transmission input shaft, on the secondary flywheel (in double-row spring sets - outer damper + inner damper) / hub disc (in single-row spring sets - only outer damper).

The springs (here F1 to F5, number basically any) can be implemented as cylinder springs, bow springs or a combination of both with different or the same spring stiffness, with or without inner spring (s). In the case shown, the springs F1 and F5 on the spring plate on each side, over catch hooks on the hub disc, caught radially, so that no Cogging / torque-related friction between these springs (F1 and F5) and the sliding track is created.

This results in a maximum of four friction points between the sliding shoes and the slide track (= MR1 to MR4) in this spring set shown as an example.

The friction-causing normal force on the sliding shoes comes on the one hand from the centrifugal force (use in rotating systems) and on the other part of the transmitted torque, since in the real case, the springs perform their spring work in a rotationally symmetric torus and part of the introduced torque as a radial force in addition acts on the sliding blocks ( 3 ).

In 4 is the case where only the spring 3 Doing spring work. This is z. B. then the case when the springs 1 and 2 due to lower spring stiffness already on strike and the spring 4 and 5 due to the higher rigidity (almost) do no work. This can only the spring 3 work. In this case, only the friction points MR3 and MR4 perform friction work, because only between the sliding shoes 3 / 4 and the slideway of the primary flywheel is given a relative movement, the spring stiffness / work of the spring 3 influenced, ie spring stiffness increased and reduced spring work.

An alternative control is in the 5 and 6 shown.

In 5 the spring arrangement is chosen so that only the spring 1 Spring work can be done and all other springs are either already on strike (have a lower stiffness) or do not do any work (have a higher rigidity). Before the spring 1 Work can be done first, the friction (static friction / sliding friction) of the friction points MR1 to MR4 must be overcome. This arrangement would be particularly advantageous in the power split in the event that the spring 1 the stiffest of the inserted springs ( 1 - 5 ) would.

In 6 is contrary to 5 only the spring 5 able to perform spring work. Because the spring 5 caught radially, this spring even works without friction. This circuit / arrangement of the springs would then be particularly advantageous for the power split again when the spring 5 the softest of the inserted springs would be. This could in particular the low speed range in which the excitation is usually highest, with the maximum available spring work decoupled (= reduction of rotational irregularities of the engine). All other (frictional) springs would contribute to decoupling only at higher speeds / moments.

In 7 the findings on the influence of the sequence of spring stiffness are summarized using the example of an external damper for the power split.

In the illustrated case, five individual springs are assembled with different stiffness to a characteristic.

Due to the peculiarity of the power split with two torque paths, a frictionless / low-friction design of the first spring stages (here: c1 and c2) and the highest possible friction for the harder springs (c4 and c5) is particularly advantageous. The friction-free first springs (c1 and c2) promote a sharp phase shift due to the low friction and at the same time the position of the resonance towards lower speeds. Both are for the extinction and the formation of Auslöschpunkte, in this particularly important speed range, and thus for the decoupling of advantage. Furthermore, the higher friction of the harder springs (c4 and c5) also increases the stiffness of these springs. As a result, the decoupling of the Federsatzzweiges in the higher speed range as desired worse and the decoupling of the entire system, by superimposing the direct branch, cheaper. The reason for this is that the decoupling of the Federsatzzweiges due to the unchanged other parameters, primarily the Zwischenmassenträgheit and the coupling gear ratio, in the higher speed range is too good. Due to the deliberate deterioration at higher speeds, due to the higher friction and stiffness in the springs c4 and c5, the signal of the Federsatzzweigs, the size of the amplitude, again better fits the direct branch, whereby the decoupling of the entire system is cheaper.

3.2 Characteristic variants for the power split

To adapt the decoupling behavior over the rotational speed, the parameters characteristic curve, intermediate mass inertia and coupling gear ratio are essentially suitable.

The advantages of different spring rate characteristics will be shown below.

The influence on the decoupling by adaptation of the spring stiffnesses for different speeds / speed ranges is a very advantageous, because relatively easy to implement, possibility for the optimal design of the power split.

To change the characteristic springs with different stiffnesses are used. Thus, the decoupling by the power split to different suggestions, eg. For example Engines with different numbers of cylinders, for downspeeding applications, cylinder deactivation, etc., to be tuned.

In 8th It is shown how a sum characteristic (eg consisting of inner damper and outer damper) could basically look like the power split.

Theoretically, any desired number of spring stiffnesses (c1, c2, ci,..., Ci + x) could be provided for matching to different speed / torque ranges. Thus, both the low speed range with very soft springs, the medium speed range with harder springs and the high speed range with very hard springs can be interpreted favorably. With optimum tuning, an extinguishing point (AP1 to APi + 1) could be generated for each individual stiffness and thus an ideal extinction / decoupling over the entire speed range (nleer to nmax) can be achieved.

It is irrelevant how the individual stiffness levels are shown constructively. For example, a stepped outer damper, a stepped inner damper, or a combination of both could be implemented.

In the 9 and 10 Characteristics are shown with five or four spring stiffness, which are well suited for example for engines with 4 or more cylinders and cylinder deactivation. These motors usually have higher torques at medium speeds with medium alternating torques, each remaining within a stiffness level and thus can form an extinguishing point.

In particular, in the case of cylinder deactivation, the design is to be selected so that the engine torque remains in the shutdown mode within one / less stages z. In one (c2) or two (c2 and c3) stage (s). This allows one or more extinguishing points to be displayed in full engine operation and in shutdown mode.

Furthermore, it applies in principle that a strongly progressive characteristic of the functio / performance of the power split accommodates. At low speeds (low centrifugal friction and low engine torque), the soft springs in this range work very well (resonance speed decreases, phase jump becomes sharper, much spring work / less friction work) and at the same time the parameters of the power split can be adjusted so that Ausschöschpunkte can be represented , At higher speeds (□ maximum engine torque) the decoupling of the spring-set branch is basically too good, so that a deterioration of the decoupling in this branch, due to very stiff springs, favors the extinction with the higher amplitudes of the direct branch.

Fig. 10

For engines with very high alternating torques, z. B. from idle speed at the level of the maximum engine torque or more, also applies in principle that not multiple extinguishing points can be designed exactly because the moment oscillations do not remain within a stiffness level and thus no clearly defined parameters (stiffness) at a certain speed or a certain Moment is given. For the design of an extinguishing point, however, it is necessary that the parameters intermediate mass inertia, coupling gear ratio and spring stiffness in a limited speed / torque range or ideally point (= point / range in which an extinguishing point should lie) are constant. If this condition is not met, no extinction point can be defined exactly. In such cases, z. As in high-torque three- or two-cylinder, it is appropriate to make the interpretation of the characteristic so that a larger extinguishing area is formed. This is z. B. achieved by a characteristic curve with few or only a single stage (s) / rigidity (s) ( 11 - 12 ).

In addition, the reduction of the mass moments of inertia of the planet gears is advantageous in such application. In 13 are presented favorable combinations of characteristics and reduced moment of inertia through the use of two or three planetary gears.

For higher torque and characteristics with multiple extinction points (spring stages, stiffness) are combinations as in 14 are shown advantageous.

In the 13 ( 13.1 - 13.4 ) is exemplified the constructive implementation of a double-row (outer damper and inner damper), multi-stage characteristic shown. In 13.1 is the construction of the whole system power split shown. The 13.2. and 13.3 show the arrangement and control of the spring sets and in 13.4 the corresponding characteristic curves are shown.

In the other ( 16 - 20 ) Further example characteristics of different design are shown.

3.3 Boundary conditions in the design of series characteristics

In the design of series spring sets in addition to the decoupling function more Criteria that have not been explicitly described so far. This includes z. B. the consideration of a starting stage (if necessary - when using a starter generator or in cooperation with hybrid applications, the engine must start not necessarily on the spring set and thus not by the spring replacement resonance), the design of a stop step, the compliance of certain stiffness jumps in the step transitions (ci / ci + 1 should be <3, otherwise torque peaks in the drive train could possibly occur due to the stop of the sliding shoes and thus impair the decoupling again.) and what must always be taken into account in the spring design are the tolerable stresses in the springs.

QUOTES INCLUDE IN THE DESCRIPTION

This list of the documents listed by the applicant has been generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Cited patent literature

• DE 102012218729 A1 [0032]
• DE 102012214361 A1 [0032]
• DE 102011007118 A1 [0032]

Claims (15)

Torsional vibration damping arrangement ( 10 ) for the drive train of a motor vehicle, comprising - an input area to be driven for rotation about an axis of rotation (A) ( 50 ) and an output area ( 55 ), the entrance area ( 50 ) a primary mass ( 1 ) and the output area ( 55 ) a secondary mass ( 2 ) and - one with the output area ( 55 ) associated coupling arrangement ( 41 ), wherein the coupling arrangement ( 41 ) a first input element ( 31 ), a second input element ( 32 ) and an output element ( 33 ), and - a torque transmission path ( 46 ) for transmitting a total torque (Mges), which between the input area ( 50 ) and the exit area ( 55 ), wherein the torque transmission path ( 46 ) from the entrance area ( 50 ) to the coupling arrangement ( 41 ) in a first torque transmission path ( 47 ), for transmitting a first torque portion (Ma1), and in a parallel second torque transmission path ( 48 ), for transmitting a second torque portion (Ma2), wherein the first and the second torque transmission path ( 47 ; 48 ) and thus the first and the second torque component (Ma1, Ma2) on the coupling arrangement ( 41 ) is recombined into an output torque (mouse), and - a phase shifter assembly ( 43 ) in the first torque transmission path ( 47 ), comprising a vibration system ( 56 ) with an input element ( 70 ) and with a against a first spring arrangement ( 4 ) about the axis of rotation A rotatable output element ( 71 ), wherein the first spring arrangement ( 4 ) at least a first spring element ( 12 ), a second spring element ( 13 ) and a last spring element ( 14 ), wherein the spring elements ( 12 . 13 . 14 ) are connected in series and where - one from the entrance area ( 50 ) incoming input torsional vibration (EDSw) by passing over the first and the second torque transmission path ( 47 ; 48 ) is divided into a first torsional vibration component (DSwA1) and a second rotational vibration component (DSwA2) - and wherein during operation of the vibration system ( 56 ) in a speed range above at least one limit speed at which the vibration system ( 56 ) is operated in a resonance range, the first torsional vibration component (DSwA1) with the second torsional vibration component (DSwA2) on the coupling arrangement ( 41 ) is superimposed so that the first torsional vibration component (DSwA1) and the second torsional vibration component (DSwA2) destructively overlap and thereby at the output element ( 33 ) of the coupling arrangement ( 41 ) is present compared to the input torsional vibration (EDSw) minimized output torsional vibration (ADSw), characterized in that the first spring element ( 12 ) compared to the second spring element ( 13 ) has a lower rigidity and that the second spring element ( 13 ) compared to the last spring element ( 14 ) has a lower rigidity Torsional vibration damping arrangement ( 10 ) according to claim 1, characterized in that the coupling arrangement ( 41 ) a planetary gear ( 61 ) with a planet carrier ( 9 ), one on the planet carrier ( 9 ) planetary pin ( 65 ) and one on Planetenradbolzen ( 65 ) rotatably mounted Planetenradelement ( 45 ), wherein the planetary gear element ( 45 ) with the entrance area ( 50 ) by means of the first input element ( 31 ) and by means of the second input element ( 32 ) and wherein the planetary gear element ( 45 ) by means of the output element ( 33 ) with the exit area ( 55 ) connected is. Torsional vibration damping arrangement ( 10 ) according to claim 1 or 2, characterized in that the first spring element ( 12 ) compared to the second spring element ( 13 ) in dependence on a rotational speed of the torsional vibration damping arrangement ( 10 ) is mounted with less friction around the axis of rotation A. Torsional vibration damping arrangement ( 10 ) according to claim 3, characterized in that the first spring element ( 12 ) by means of a sliding shoe ( 7 ) against a slideway ( 51 ) is mounted radially. Torsional vibration damping arrangement ( 10 ) according to claim 4, characterized in that the slideway ( 51 ) from the input element ( 70 ) or from the starting element ( 71 ) or partially from the input element ( 70 ) and the output element ( 71 ) of the phase shifter assembly ( 43 ) is formed. Torsional vibration damping arrangement ( 10 ) according to claim 4 or 5, characterized in that the sliding shoe ( 7 ) is mounted by means of a guide element in the radial direction, so that under a centrifugal force, caused by a rotation of the torsional vibration damping arrangement ( 10 ) about the axis of rotation A, the shoe no radial force due to the centrifugal force on the slide ( 51 ) exercises. Torsional vibration damping arrangement ( 10 ) according to one of claims 1 to 6, characterized in that the second spring element 13 and / or the last spring element 14 by means of a sliding shoe ( 7 ) at the slideway ( 51 ) is stored. Torsional vibration damping arrangement ( 10 ) according to claim 7, characterized in that the sliding shoe ( 7 ) of the second spring element ( 13 ) to the slideway ( 51 ) has a lower coefficient of friction μ, than the sliding shoe ( 7 ) of the last spring element ( 14 ) to the slideway ( 51 ). Torsional vibration damping arrangement ( 10 ) according to one of claims 1 to 8, characterized in that to the first spring arrangement ( 4 ) a second spring arrangement ( 18 ) is connected in series, wherein the second spring arrangement ( 18 ) radially inwardly while axially overlapping to the first spring arrangement ( 4 ) is located. Torsional vibration damping arrangement ( 10 ) according to one of claims 1 to 8, characterized in that to the first spring arrangement ( 4 ) a second spring arrangement ( 18 ) is connected in series, wherein the second spring arrangement ( 18 ) at least partially radially overlapping and thereby axially staggered to the first spring arrangement ( 4 ) is located. Torsional vibration damping arrangement ( 10 ) according to claim 10, characterized in that the first and the second spring arrangement ( 4 . 18 ) radially outward and axially overlapping to the coupling arrangement ( 41 ) is arranged. Torsional vibration damping arrangement ( 10 ) according to one of claims 9 to 11, characterized in that an overall rigidity of the first spring arrangement ( 4 ) is smaller than an overall rigidity of the second spring arrangement ( 18 ). Torsional vibration damping arrangement ( 10 ) according to one of claims 9 to 11, characterized in that an overall rigidity of the first spring arrangement ( 4 ) is greater than an overall stiffness of the second spring arrangement ( 18 ). Torsional vibration damping arrangement ( 10 ) according to one of claims 9 to 13, characterized in that the second spring arrangement ( 18 ) at least one spring element ( 15 ), wherein the spring element ( 15 ) is a cylinder pressure spring. Torsional vibration damping arrangement ( 10 ) according to one of claims 1 to 13, characterized in that the spring elements ( 12 . 13 . 14 ) of the first spring arrangement ( 4 ) Are cylinder compression springs.

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