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WO2012119751A1 - Volant bi-masse monté dans la chaîne cinématique d'un véhicule à moteur - Google Patents

Volant bi-masse monté dans la chaîne cinématique d'un véhicule à moteur Download PDF

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Publication number
WO2012119751A1
WO2012119751A1 PCT/EP2012/000977 EP2012000977W WO2012119751A1 WO 2012119751 A1 WO2012119751 A1 WO 2012119751A1 EP 2012000977 W EP2012000977 W EP 2012000977W WO 2012119751 A1 WO2012119751 A1 WO 2012119751A1
Authority
WO
WIPO (PCT)
Prior art keywords
spring unit
spring
dual
secondary part
mass flywheel
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Ceased
Application number
PCT/EP2012/000977
Other languages
German (de)
English (en)
Inventor
Herbert Meyer
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Audi AG
Original Assignee
Audi AG
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Audi AG filed Critical Audi AG
Publication of WO2012119751A1 publication Critical patent/WO2012119751A1/fr
Anticipated expiration legal-status Critical
Ceased legal-status Critical Current

Links

Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F15/00—Suppression of vibrations in systems; Means or arrangements for avoiding or reducing out-of-balance forces, e.g. due to motion
    • F16F15/10—Suppression of vibrations in rotating systems by making use of members moving with the system
    • F16F15/12—Suppression of vibrations in rotating systems by making use of members moving with the system using elastic members or friction-damping members, e.g. between a rotating shaft and a gyratory mass mounted thereon
    • F16F15/131—Suppression of vibrations in rotating systems by making use of members moving with the system using elastic members or friction-damping members, e.g. between a rotating shaft and a gyratory mass mounted thereon the rotating system comprising two or more gyratory masses
    • F16F15/133—Suppression of vibrations in rotating systems by making use of members moving with the system using elastic members or friction-damping members, e.g. between a rotating shaft and a gyratory mass mounted thereon the rotating system comprising two or more gyratory masses using springs as elastic members, e.g. metallic springs
    • F16F15/134—Wound springs
    • F16F15/13469—Combinations of dampers, e.g. with multiple plates, multiple spring sets, i.e. complex configurations
    • F16F15/13476—Combinations of dampers, e.g. with multiple plates, multiple spring sets, i.e. complex configurations resulting in a staged spring characteristic, e.g. with multiple intermediate plates

Definitions

  • the invention relates to a arranged in the drive train of a motor vehicle two-mass flywheel according to the preamble of claim 1.
  • Dual mass flywheels (DMF) as torsional vibration dampers in the drive train of motor vehicles are widely known.
  • the interpretation of the Torsionskennline determining spring unit (usually annularly arranged helical compression springs) of the DMF is such that its resonant frequency is outside the regular engine operation of idle speed to full load speed to ensure an always effective torsional vibration damping.
  • the resonant frequency may be well below the idle speed at an engine speed, so that this condition is passed through only briefly when starting the engine or when it is stopped.
  • the torsional characteristics of dual-mass flywheels today are usually designed so that the alternating torque of the engine from idle speed to maximum speed and the maximum thrust torque is transmitted to the maximum torque torsionally soft. This design leads to the known resonant frequency of the two-mass oscillator. The resonance frequency may not be excited by the combustion engine, since the resulting vibration amplitudes
  • Today common DMFs for 4-cylinder engines can have a resonant frequency in the range of 12Hz, so this at about 350rpm. Engine speed is hit. This value then increases to 700 rpm in 2-cylinder operation. Since it is a damped vibration system, the increase in amplitude in the range of the resonance against the exciting amplitude of the motor is up to 1400rpm. so high that at least loss of comfort in terms of roar and hum of the body arise. Therefore, operation with only 2 cylinders is only permitted above this speed. The CO2 savings potential is therefore clearly limited.
  • a ZMS is known in which by means of an electrically or hydraulically operated adjusting device, the spring unit or at least one spring of the spring unit for varying the resonant frequency of the ZMS acted upon or is adjustable in its bias. In particular, this should improve the vibration damping when starting the engine and during its run-up.
  • the object of the invention is to form the Torsionsdämpferkennline arranged in the drive train of a motor vehicle ZMS with structurally simple means so that the excitation of its resonant frequency can no longer take place at any moment point so that disturbing or component hazardous vibrations can occur.
  • a arranged in the drive train of a motor vehicle ZMS with structurally simple Means are designed so that the operation of the internal combustion engine in the lower load range is also possible near its resonance frequency.
  • the spring unit is arranged biased between the primary part and the secondary part, that it produces a rigid connection below a defined engine torque and that further the spring unit is driven downstream of a second spring unit, which is essentially effective only below the defined engine torque ,
  • the ZMS is limited in its effect on the load range above the selected biasing torque.
  • the task of vibration isolation of the drive train up to this biasing moment is transmitted to the ZMS downstream torsion damper.
  • the ZMS behaves almost rigidly.
  • the vibration decoupling is done completely in the downstream torsion damper. To split the effect of the two torsional damper no actuator or the like is necessary.
  • the characteristic curve of the dual-mass flywheel is therefore modified to such an extent by the installation of the torsion springs in the strongly preloaded state that a movement in the DMF takes place only from a freely definable engine torque.
  • This moment can be, for example, the moment to which a 2-cylinder operation is intended.
  • the R4 Operation resumed and the ZMS works as known with its intended characteristic.
  • a torsion damper is provided after the DMF whose characteristic curve extends at least up to the preloaded moment in the DMF.
  • the usual torsional damper of clutch plates for powertrains with Einmassenschwungrad offers. Since only low torques have to be covered, a very good decoupling for all low load ranges in push and pull operation can be achieved here as well. A damper for decoupling in idle can also be provided.
  • the abutment torque of the second torsional damper must be selected slightly above the breakaway torque of the first torsional damper.
  • This method is of course not limited to the 2-cylinder operation of a 4-cylinder engine, but can be applied to all known engine designs and suggestions by more or less cylinder at the shutdown.
  • the transient torque of the EMS in a DMF must then be adjusted (as well as the downstream torsion damper). Due to the bias of the springs, the ZMS becomes EMS until a certain moment.
  • a two-cylinder operation is possible up to low speeds.
  • the coupling of the secondary side increases the mass moment of inertia of the flywheel and thus reduces the alternating torque delivered into the drive train.
  • the effect of the downstream torsional damper is thereby improved.
  • the effect of a built-in or on the ZMS order styling device such as a centrifugal pendulum is not hindered.
  • the consumption-saving potential can be fully exploited.
  • the second spring unit can be switched on in an advantageous manner between the first secondary part interacting with the primary part and a second secondary part.
  • the second secondary part can already be a driven-side gear part, such as a correspondingly modified input flange, etc.
  • the second spring damper unit which acts below the defined engine torque, can be formed by the spring unit in the driver disk with virtually no additional structural expenditure. This then only requires a suitable adaptation of the two spring units in the above sense.
  • the two spring units can be formed in a manner known per se by means of compression springs arranged annularly around the primary part, the secondary part and optionally the driver disk, between which the each abortive driver of the secondary parts are clamped backlash, the spring rate of the springs is tuned accordingly.
  • the second spring unit is designed for the resulting lower engine torque and the first spring unit to the higher engine torque during operation of the internal combustion engine with all cylinders.
  • the second spring unit may be tuned to accommodate the changed vibration excitation throughout the low load RPM range.
  • the internal combustion engine can also be operated at low speeds without reaching the critical range of the resonance frequency of the ZMS.
  • Fig. 2 shows another vibration model of a ZMS with a
  • Torsionskennline which only allows a rotation from a certain torque, with a downstream torsion damper, whose torsional characteristic just above this torque mentioned no longer allows the further rotation.
  • FIG. 1 schematically shows the vibration model of a dual-mass flywheel 10 (DMF) which is switched on in a drive train of a motor vehicle as a torsional vibration damper whose primary part 12, which is fixedly connected to the crankshaft, not shown, of a reciprocating internal combustion engine has a torsionally elastic limited elasticity via a spring unit 14 and a damper unit 16 a secondary part 18 is coupled.
  • DMF dual-mass flywheel 10
  • the secondary part 18 is connected for example via an integrated friction clutch (not shown) and a transmission input shaft 19 (acting as a torsion spring) with the speed change gear 20 and via this with the other drive train 21 of the motor vehicle 22, which components complement the illustrated vibration model accordingly ,
  • the spring unit 14 with the damper unit 16 is vibrationally tuned so that the drawn in the graph below torsion (torque over angle of rotation) 24 increases linearly with increasing motor torque M above the swing angle phi, for example. Up to the stop moment Mmax, the DMF is effective as a torsional vibration damper.
  • Fig. 2 shows a modified to Fig. 1 vibration model in which between the primary part 12 of the DMF 10 ' and the secondary part 18 a parallel to the damper unit 6, the first spring unit 26 is turned on.
  • the spring unit 26 is composed of several, biased between the primary part 12 and the secondary part 18 in the circumferential direction backlash built screw compression springs together.
  • the bias of the compression springs of the spring unit 26 is designed so that the rotational connection between the Primary part 12 and the secondary part 18 below a defined engine torque M (is still running) is rigid, so no torsional vibration damping takes place.
  • the secondary part 18, a second secondary part 28 is driven downstream, which is connected via a second spring unit 30 and also a damper unit 32 to the first secondary part 18 is limited torsionally elastic coupled.
  • the second spring unit 30 lower spring rate is designed so that it acts as a vibration isolation at lower engine torque.
  • the graph drawn below the vibration model according to FIG. 2 shows the torsion characteristic 24 'of the DMF 10 ' and the torsion characteristic curve 34a, 34b of the downstream torsion damper 11, which is composed as follows:
  • the spring unit 30 acts with the damper unit 32 (characteristic 34 a, 34 b) vibration-isolating.
  • the region 34a of the characteristic curve represents, for example, the very flat spring stage required for vibration decoupling during idling and neutral gear.
  • the area 34b shows the position of the stop torque of the torsion damper 11 just above the engine torque Mi.
  • the first spring unit 26 is set rigidly up to this moment Mi by bias and allows no relative movement between the primary part 12 and the first secondary part 18 to. With increasing moment M, the first spring unit 26 with the damper unit 16 becomes active and the torsional vibration damping now takes place between the primary part 12 and the secondary part 18.
  • the spring unit 26 has the same spring rate as the spring unit 14 of FIG. 1.
  • the biasing torque of the DMF 10 ' is chosen so that the low torque in two-cylinder operation just does not lead to a movement of the torsion damper.
  • the high alternating torques generated by the gas force also apply to the prestressed torsion damper and the first secondary part, the bias voltage increases the possible approach to the resonance frequency.
  • This design advantageously takes into account the changed excitation frequency of the engine when the cylinders are switched off or in the two-cylinder mode.
  • the bias prevents the ZMS resonance, which is very close to the speed range selectable by the driver, to be excited.
  • the drive train is decoupled in this operating state by the second torsional damper 11 of disturbing vibrations. With increasing loads through the connection of the cylinder and the excitation frequency increases again and the resonance point is again far below the approachable speed range.
  • the second secondary part 28 of the torsion damper 11 may be in a manner not shown in an embodiment with integrated friction clutch whose drive plate, in which the second spring unit 30 may be integrated in a conventional manner. This is just the well-known To modify spring unit to a suitably tuned spring unit 30.
  • the second secondary part 28 and its spring and damper unit 30, 32 may be arranged, for example, in an input flange of the downstream variable speed transmission 20.
  • the spring unit 26 may be formed in a conventional manner by circularly about the rotationally symmetrical primary part 12 and secondary part 18 arranged helical compression springs which are inserted into corresponding recesses of the primary part 12 and engage in the backlash and under the defined bias driver of the secondary part 18, for example in the aforementioned DE 10 2007 044 474 A1 in FIG. 1.
  • the spring unit 30 may be formed as known per se in driver disks of friction clutches also by circularly arranged helical compression springs, which are matched in their spring hardness corresponding to the spring unit 30.
  • the damper units 16 and 32 may be in a known manner friction damper, as are common in two-mass flywheels or Mitêtn of friction clutches.
  • the invention is not limited to the embodiment shown in the vibration model Fig. 2.
  • the second spring unit only has to be arranged at one point between the DMF and the gear mass.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Acoustics & Sound (AREA)
  • Aviation & Aerospace Engineering (AREA)
  • Mechanical Engineering (AREA)
  • Vibration Prevention Devices (AREA)
  • Mechanical Operated Clutches (AREA)

Abstract

L'invention concerne un volant bi-masse (ZMS) monté dans la chaîne cinématique d'un véhicule à moteur, comprenant une partie primaire symétrique en rotation, qui est reliée au vilebrequin d'un moteur à combustion interne, une partie secondaire qui est en sortie directe ou indirecte sur un arbre de transmission, et une unité élastique d'amortissement qui absorbe les vibrations torsionnelles émanant du vilebrequin, la fréquence de résonance du volant bi-masse étant variable. L'invention vise à proposer une conception de ZMS favorisant son adaptation aux différentes conditions de fonctionnement. A cet effet, l'unité élastique (26) est disposée de manière précontrainte entre la partie primaire (12) et la partie secondaire (18) de sorte que, en-dessous d'un couple moteur défini, elle crée une liaison rigide, l'unité élastique (26) étant en outre, en termes d'entraînement, en aval d'une deuxième unité élastique (30) qui n'agit sensiblement qu'en-dessous du couple moteur défini.
PCT/EP2012/000977 2011-03-10 2012-03-05 Volant bi-masse monté dans la chaîne cinématique d'un véhicule à moteur Ceased WO2012119751A1 (fr)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE102011013483.2 2011-03-10
DE102011013483.2A DE102011013483B4 (de) 2011-03-10 2011-03-10 Im Antriebsstrang eines Kraftfahrzeuges angeordnetes Zweimassenschwungrad

Publications (1)

Publication Number Publication Date
WO2012119751A1 true WO2012119751A1 (fr) 2012-09-13

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Application Number Title Priority Date Filing Date
PCT/EP2012/000977 Ceased WO2012119751A1 (fr) 2011-03-10 2012-03-05 Volant bi-masse monté dans la chaîne cinématique d'un véhicule à moteur

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DE (1) DE102011013483B4 (fr)
WO (1) WO2012119751A1 (fr)

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP2017511271A (ja) * 2014-03-19 2017-04-20 ツェットエフ、フリードリッヒスハーフェン、アクチエンゲゼルシャフトＺｆ Ｆｒｉｅｄｒｉｃｈｓｈａｆｅｎ Ａｇ ハイブリッドモジュール及びハイブリッドモジュールを有するドライブトレーン

Families Citing this family (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2013083106A1 (fr) * 2011-12-05 2013-06-13 Schaeffler Technologies AG & Co. KG Chaîne cinématique
DE102014210449A1 (de) * 2014-06-03 2015-12-03 Schaeffler Technologies AG & Co. KG Torsionsschwingungsdämpfer, Verfahren zum Auslegen eines Torsionsschwingungsdämpfers,sowie Drehmomentübertragungseinrichtung
DE102014218926A1 (de) * 2014-09-19 2016-03-24 Zf Friedrichshafen Ag Drehschwingungsdämpfer und Anfahrelement
DE102014220927A1 (de) * 2014-10-15 2016-04-21 Schaeffler Technologies AG & Co. KG Drehschwingungsdämpfer
DE102016208261A1 (de) * 2016-05-13 2017-11-16 Schaeffler Technologies AG & Co. KG Drehschwingungsdämpfer, insbesondere für einen Drehmomentwandler und Drehmomentwandler mit diesem

Citations (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102007044474A1 (de) 2007-09-18 2009-03-19 Magna Powertrain Ag & Co Kg Zweimassenschwungrad
WO2010043301A1 (fr) * 2008-10-17 2010-04-22 Luk Lamellen Und Kupplungsbau Beteiligungs Kg Amortisseur de torsion à deux voies
DE102009013965A1 (de) * 2009-03-19 2010-09-23 Daimler Ag Dämpfungseinrichtung

Family Cites Families (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102004024747A1 (de) * 2004-05-19 2005-12-15 Zf Friedrichshafen Ag Torsionsschwingungsdämpfer
DE102007049075A1 (de) * 2007-10-12 2009-04-16 Zf Friedrichshafen Ag Torisionsschwingungsdämpferanordnung

Patent Citations (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE102007044474A1 (de) 2007-09-18 2009-03-19 Magna Powertrain Ag & Co Kg Zweimassenschwungrad
WO2010043301A1 (fr) * 2008-10-17 2010-04-22 Luk Lamellen Und Kupplungsbau Beteiligungs Kg Amortisseur de torsion à deux voies
DE102009013965A1 (de) * 2009-03-19 2010-09-23 Daimler Ag Dämpfungseinrichtung

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP2017511271A (ja) * 2014-03-19 2017-04-20 ツェットエフ、フリードリッヒスハーフェン、アクチエンゲゼルシャフトＺｆ Ｆｒｉｅｄｒｉｃｈｓｈａｆｅｎ Ａｇ ハイブリッドモジュール及びハイブリッドモジュールを有するドライブトレーン

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Publication number Publication date
DE102011013483A1 (de) 2012-09-13
DE102011013483B4 (de) 2015-12-10

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