JPS6135416B2 - - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION The present invention relates to a torsional vibration damper of the type having a hub attached to an outer inertial part by an elastomeric ring.

The invention is particularly useful for damping torsional vibrations in internal combustion engines. Such dampers are generally currently classified in the United States Patent Office as Class 74 subclass 574. Torsional vibration can be thought of as a back-and-forth twist of the crankshaft that is added to the primary unidirectional rotation of the crankshaft of an internal combustion engine. If uncontrolled, such torsional vibrations can cause damage to other parts of the engine or its cooling system, particularly where one of the resonant frequency modes of the crankshaft coincides with a particular ignition frequency of the engine or with a particular harmonic of that frequency. as well as contributing to damage to the crankshaft. According to the current theory of elastomeric vibration dampers, the torsional vibrational energy transferred to the crankshaft by the action of the piston is converted into heat within the elastomer. The elastomer can therefore be thought of as a drain or sump that constantly receives a portion of the energy that produces torsional vibrations.

Such damping device configurations commonly include an outer or annular member of ring or annular shape of some significant mass. The inner portion of the ring is attached to an elastomeric ring that is attached to a hub or other element that is attached to a rotating crankshaft of the engine. Both the hub and the annular member may be cast iron. As the crankshaft rotates, each increase in torque causes a slight acceleration in the metal adjacent the crank arm, such as due to the rapid combustion of fuel within the cylinder. Due to the elastic or resilient properties of the metal, when the metal compensates, it rotates slightly in the opposite direction. Such pressure causes torsional vibrations in the shaft. A typical example of a torsional vibration is an angular vibration at a frequency of 150 to 250 cycles per second, with an amplitude of 1/4 to 1 degree, accompanied by an engine crankshaft rotating at 3000 rpm.

The purpose of a torsional vibration damper is to reduce the amplitude of torsional vibrations. Such a reduction reduces the required strength of the crankshaft, thereby reducing the weight of the crankshaft. The damper directly affects the crankshaft and also prevents vibrations of various other components of the internal combustion engine caused by the vibrations of the crankshaft.

When an internal combustion engine operates at various engine speeds,
Several vibration frequencies appear in the crankshaft. In general, most automotive and diesel engines that do not utilize current designs of torsional vibration dampers have resonant frequencies of fairly large amplitude within the engine operating speed range. However, at any given engine speed, there are torsional vibrations from various vibrations, and these can be large.

Dry friction to dampen vibrations
frition) is well known in the art. It is known to use dry friction alone as a damping means and in parallel with viscous damping means, which can be utilized by distorting elastomer-type materials. Vibration damper connection〓〓〓〓
In terms and terminology, in a parallel combination of damping means, the forces or torques transmitted through the two means are unequal, and in a series combination of damping means, the forces or torques transmitted through the two means are unequal. The time average direction of torque is always the same. Examples of US patents that include dry friction only as a damping means include: MacPherson No.
No. 1670369, Meyer No. 1967446, Meyer No. 2062369, Breer No.
No. 1840655, No. 1840656 of Braille, No. 2960189 of Os burn. The three main drawbacks of pure dry friction dampers are: (1) Due to the mechanical requirements for the bearing surfaces between the inertia member and the hub, the radial unbalance of the high speed rotating part The difficulty of maintaining radial concentricity between the inertia member and the hub required to minimize (2) Additional balance problems due to the fact that the inertial mass can change its angular position relative to the hub, and therefore the balance of all parts can change constantly during operation. (3) The amount of relative vibration amplitude is limited by the pure dry frictional contact between the inertial mass and the hub. The amount of energy absorbed by such a device is a function of the relative motion between the two principal members, which in this case is limited to the magnitude of the vibration of the hub, and It only vibrates at very high frequencies relative to the design frequency. Examples of US patents for dry friction used in parallel with elastomeric viscous friction include: Griswold No. 1984579; Griswold No. 1984579;
No. 2083561, Christman No.
No. 2153194, McFarland No.
No. 2383400, Kearns No. 2440956
No. 3020739 of Roe, No. 3075405 of Krebs, and No. 3077123 of Katzenberger. Applying elastomeric and dry friction damping means in parallel,
Although it effectively solves the unbalance problem of pure friction damper, it has the drawback of weak low frequency response. At a predetermined hub vibration amplitude and at a frequency below a value that causes acceleration of the inertia ring of a significant magnitude to exceed the static or destructive torque of the contacting dry friction surface, the inertia ring is locked to the hub; There is no relative movement between the inertia ring and the hub, so there is no damping and the damper is essentially a dead mass. Therefore, in a parallel elastomeric dry friction damper operating at below-destructive acceleration conditions, all inertia ring torque is transmitted through the dry friction means and is transferred through the elastomeric means. There is no torque transmission. Hann, U.S. Pat. No. 1,913,984 is an example of a pure series combination of elastomeric and dry friction damping means in which the force produced by inertial mass vibrations transmitted through the dry friction means increases over time. The average deflection is equal to the force transmitted through the elastomer means. The above U.S. patent no.
In pure series operation, such as in 1913,984, essentially a pure dry friction damper acts in series with the elastomeric means and all the problems of pure dry friction dampers remain.

It is known in the art to use two vibratory or inertial masses, each with a different mass, to attenuate two specific frequencies. A parallel damper arrangement can therefore best damp the two main resonant frequencies. Multi-mass dampers can similarly damp multiple resonant frequency vibrations. A known parallel damper is the Krebs U.S. Patent No.
No. 3,075,405 and FIG. 2 of Pierce US Pat. No. 2,477,081 are known, each of which has a separate and independent It is attached to the hub by an elastomeric member.

The aforementioned US Pat. No. 1,967,446 shows two masses designed to use dry friction as a damping means. In general, known techniques using one or more inertial masses in vibration dampers utilize purely elastomeric or dry friction means, or a parallel combination of two means, with the masses acting essentially independently of each other. is used.

According to the practice of the invention, elastomer and dry damping means are used together to create a vibration damper. The elastomeric and dry damping means act in combination in series and this combination acts in parallel with the elastomeric means. The series combination forms a non-linear spring and when used in parallel with the elastomeric means creates a damper with the desired predetermined non-linear dynamic characteristics when vibrations are applied.
provide. This combination removes the design limitations of each, while retaining the desirable characteristics of a pure elastomer damper and pure dry friction. The result is an extra damping capability that is better than using either means alone. In other specific examples,
Multiple inertial masses are mounted and supported on the same elastomeric member. The entire damper is assembled in one step and utilizes elastomeric members with geometric cross-sections to create axial connectivity.
Series rubber dry friction means are used between the masses to take advantage of relative motion between the individual masses as well as relative motion between the inertial masses and commonly utilized hub members. In other embodiments, multiple inertial masses are similarly mounted and supported on the same elastomeric member, and axially variable radial deformation of the rubber is used to provide the case where there is no interaction. any desired degree of dry friction mass interaction. In other embodiments, dry friction is used in parallel with the elastomeric damping means to provide greater axial cohesion. In general, the present invention relates to the use of dry friction means and/or elastomeric damping means, combining single or multiple inertial masses to create a vibration damper with a predetermined nonlinearity; The present invention relates to the ability to operate over a wider range of effective frequencies than would otherwise be possible, or even if both measures were used simply in parallel.

Next, a torsional vibration damper according to the present invention will be described in detail with reference to the accompanying drawings.

First of all, referring to FIG. 1, reference number 1
0 generally indicates a torsional vibration damper constructed according to the invention, the torsional vibration damper 10 being designated by the reference numeral 12 and adapted to be connected to a shaft (not shown) having an axis of rotation 14. It has a built-in hub. The outer radial surface of the hub is designated by reference numeral 16 and the hub supports a continuous annular elastomer band designated by reference numeral 18. Reference numeral 20 is a first inertia ring (inertia member) having an innermost radial surface 22 in contact with the elastomeric band. A second inertia ring is also supported by an elastomer band 18, which inertia ring has the reference numeral 24.
and has an innermost radial surface 26. Reference numeral 30 indicates a rounded chamfer or assembly radius near the left axial surface of ring 20, and reference numeral 32 indicates a similar chamfer near the left axial surface of the hub. Similar chamfered assembly radii 34 and 86 are also provided on the right axial face of the torsional vibration damper. Reference numerals 38 and 40 indicate additional assembly radii of the innermost radial portions of the contact surfaces of the inertia rings 20 and 24. A continuous angular groove 41 is formed by them and a corresponding groove 42 located axially in the groove 41
is provided on the hub member. It will be appreciated that each of the several assembly radii is annular and continuous.

The elastomer 18 is in a radially deformed (compressed) state and therefore has some bulge at its two axial ends. The elastomer is typically adhesively bonded to both the hub member and the inertia ring.
Adhesive techniques for externally supporting the inertia ring are well known in the art and do not form part of this invention.

The operation of the damper illustrated in FIG. 1 is as follows. Torsional vibrations from the internal combustion engine are transmitted via the crankshaft to the hub 12, whose axis of rotation 14 often coincides with the axis of rotation of the crankshaft. The hub 12 undergoes torsional oscillations and the cast iron inertia rings 20 and 24 are retarded or out of phase with respect to the hub 12 due to the resiliency of the elastomeric material 18. In a manner well known to those skilled in the art, the elastomeric member 18 converts at least a portion of the energy associated with the torsional vibrations into heat, thereby preventing the amplitude from reaching dangerously high levels.

In the inner region of the annular surface 22 of the ring 20, the elastomer is deformed radially, while in the left axial surface and grooves 41, 42, the elastomer is free and extended so that there is little or no radial deformation. do not have. Similarly, radial deformation of the elastomer is caused by inertia ring 24 except at the axial end of inertia ring 24.
along the inner surface 26 of 4. The radial deformation therefore creates radial stresses on the two inertia rings, except in the immediate vicinity of the axial ends.

The particular geometry resulting from the conventional grooves 41 and 42 is not limited to being formed by various or single arcs as shown in FIG. In practice, these grooves are relatively easy to machine.
With this in mind, it can be formed into a wide variety of shapes. The grooves may be geometrically similar, such as 62 and 64 in FIG. 2, or 41 and 42 in FIG.
They may be dissimilar, such as. The limiting criterion for the grooves is simply that the rubber has approximately zero radial compression in the axial vicinity of the intermediate plane between the various inertial members. By substantially zeroing the radial rubber compression in the axial vicinity of the intermediate plane between the masses or inertia rings, a force is created in the rubber that is exerted on the inertia rings that would normally tend to pull the rings apart. The ring does not move axially by any significant amount due to the frictional or adhesive restraint of the elastomeric members, resulting in very small or almost zero normal forces between the inertial members. The general design criteria for multi-mass vibration dampers is that each of the masses has a specific resonant frequency that is different from the resonant frequency of the pond mass, and each of the masses is generally out of phase with every other mass. This means that there is a relative back-and-forth movement between the active masses of the damper. what if,
If a normal force exists between the masses, this relative motion creates a frictional drag between the masses such that each mass impedes the motion of any adjacent mass. Therefore, heat is generated at the intermediate plane between the masses and vibrational energy is dissipated. Although this mass interaction is not always a desirable property in multi-mass dampers, achieving the desired mass interaction is a function of the system in which the damper is used. fact,
It is a function of the embodiment of FIG. 1 to minimize this mass interaction by radial compression of the elastomer near the intermediate plane between the masses. Therefore, the first
The illustrated embodiment produces a near-zero mass interaction and exhibits a dynamic behavior very similar to two separate dampers with different resonant frequencies. Differences in resonant frequencies are caused by differences in rotational inertia between the masses and/or by varying the rotational spring rate of the elastomeric member by varying the axial width of the portion of the elastomeric member that directly cooperates with a given mass. . Zero mass interaction between the masses of a damper using multiple masses or inertia rings and a single common elastomeric material can also be achieved in the manner shown in FIGS. 9 and 10. In FIG. 9, a peripheral groove is cut in the intermediate plane between the masses and there is no attempt to alter the radial compression at the intermediate plane of the rubber. In that way, the bearing surface between the masses is minimized and
Frictional drag or mass interactions are also minimized. Zero mass interaction can be achieved in multi-mass dampers utilizing a single elastic member, such as that shown in FIG. 10, where a thin spatial element is provided during assembly and later separated from the mass. This can be done either by heating the spacer (such as a metal washer) before assembly and allowing the element to cool and contract axially after assembly, eliminating contact between the masses, or by soft graphing. This can also be achieved by making the spacer of any convenient material, such as an ait washer, which is quickly worn away by the action of relative movement between the masses. Each of the two inertia rings 20 and 24 has the function of damping certain torsional vibration frequencies well known to those skilled in the art and described, for example, in the above-referenced U.S. patents to Klebes and Puerez.

Next, a description will be given with reference to FIG. This example is a similar example to the previous example, except that there are four inertia rings instead of two. Reference numeral 50 designates a hub rotatable about axis of rotation 14 , the outer radial surface of which supports an elastomer band 52 . Inertia ring 5
4, 56, 58 and 60 are attached to a common elastomer band 52, and the inner radial surfaces of these inertia rings are preferably bonded to the elastomer such that the elastomer is bonded to the hub. Assembly radii are again used, one such radii being shown at 62 on the left innermost radial surface of the inertia ring 54. Similar radii are formed on the other inertia rings as shown. Groove 6
An annular groove such as 4 is also provided in the outer radial surface of the hub 50 at a corresponding axial location, thereby providing a groove in which the radially deformed or compressed elastomer 52 is freed and filled. Form. At both axial ends of the damper, there is no radial compression of the elastomer, as indicated by the protrusions. Similarly, there is little or no radial deformation in the continuous annular grooves corresponding to 62 and 64, as well as the grooves not indicated by reference numerals. Similarly, the inertia ring is adapted to attenuate certain frequencies. The specific example shown in Fig. 2 is also designed so that the mass interaction is close to zero, and as explained in the two mass system, this end
Additional concepts used for purposes also apply.

Next, referring to FIG. 3, another example similar to the example shown in FIG. 1 is illustrated. Reference numeral 70 designates a hub adapted to be attached to the crankshaft of an internal combustion engine. The radially outermost periphery (radial surface) of the hub is designated by reference numeral 74 and is convex in cross-section. The inertia ring members are indicated at 20' and 24' and have surfaces 76 and 78 as shown on the inner radial portion. Reference numeral 80 indicates an elastomer band that is preferably adhered to both the hub and the inertia ring. Reference numeral 82 designates the contact axial intermediate surface of the two inertia rings. This embodiment differs from the embodiment shown in FIG. 1 in that the curved shapes of surfaces 74, 76, and 78 are utilized to axially shift the deformation of the elastomer and reduce the mutually normal forces of contact surfaces 82. The difference is that they change. Accordingly, the embodiment of FIG. 3 is designed to provide mass interaction, which is sometimes desirable for certain systems requiring dampers. The geometry of intermediate surfaces 74, 76 and 78 may vary from that shown in FIG. Issues to consider in design are as follows.
The junction between surfaces 76 and 78 should form a smooth transition in the annular cavity occupied by the resilient member after assembly. The radial compression of the elastic member after assembly should be maximum at the axial position defining the intermediate plane between the masses and should thus locally decrease in both axes from this position. .

Next, referring to FIG. 4, although only a single inertia ring is utilized, not only the elastomer converts energy into heat, but also the abrasive friction material converts energy into heat. An example is shown in which the damping effect is achieved by this. In this technical field, this latter effect is
This is called dry friction. Reference numeral 90 designates a web portion of hub 92, which
2 is connected to the rotating shaft as described above. Reference numeral 94 designates a portion of the hub 92 having a larger radius than the other portions, the radially outermost circumference of the hub being defined by a generally cylindrical portion that merges with a conical portion. Reference number 96 is
The inertia ring is shown preferably bonded at the innermost radial portion to the elastomeric band 98, which is also preferably bonded to the outermost radial portion of the hub. Since the elastomer is radially compressed, it has some bulge at the ends as shown. Reference numeral 100 indicates a threaded fastener 102 provided with multiple angular spaces, only one illustrated.
indicates a flat ring or plate attached to the left side of the hub. The outermost radial portion of the plate 100 presses against an annular continuous elastomer 104, also in the shape of a flat ring or disk;
The annular continuous elastomer 104 is pressed against a wear friction material such as a flat ring-shaped common brake fabric for automobile brake drums. Reference number 108 indicates elastomer 104 and friction material 10
6 shows an annular recess in which 6 is located. Elastomer disk 104 and friction material disk 106 form a nonlinear coupling for rotational motion.

During operation, inertia ring 96 cooperates with the hub and elastomeric member to damp vibrations in the manner described above.
When such relative motion becomes large enough,
Slippage occurs between the friction material 106 and the inertia ring. This, in turn, causes dry friction damping. No adhesive layer is required between elastomer 104 and ring 100 or friction material 106. This means that the coefficient of friction between the (metallic) inertia ring 96 and the friction material is three intermediate surfaces 100-104, 104-10.
This is because it is the smallest among the friction coefficients of 6, 106-96. In this manner, additional energy of torsional vibrations is transferred to the friction material 106 and the inertia ring 96.
is additionally converted into heat by the friction between the
The outermost radial portion of ring 100 cooperates with inertia ring 96 to form flat rings 104 and 10.
6 and compress.

The torsion spring formed by the composition of members 104 and 106 is intentionally designed to have non-linear characteristics. That is, the torque transmitted through this spring has a strain in the spring that varies directly linearly with the torque (or between the intermediate planes between 100 and 104 and between 106 and 96). (relative angular movement between the two). The behavior is shown in FIG. Linear torsion springs typically have torque as shown by curve A with the slope or first derivative of curve A representing the so-called spring rate, typically having units of inch pounds per radian. 〓
represents the displacement characteristics. A spring such as formed by the composite of members 104 and 106 exhibits characteristics similar to those shown as curve B in FIG. As the torque applied to the spring increases, the angular displacement between the axial ends of the spring increases in a manner linearly proportional to the torque as it releases the friction surfaces and begins to slip, as shown at D in FIG. . Slip continues as the torque no longer increases and the torque-displacement characteristic takes the form of the right-hand part of curve B parallel to the displacement axis. Since the spring rate or slope of curve B is not constant,
The spring defined by curve B is said to be non-linear.
The damper defined in FIG. 4 has an additional torsion spring formed as an elastomeric member 98. The springs formed by spring 98 and composites 104 and 106 act in a parallel fashion so that all spring rates resisting movement of inertial member 96 relative to hub 92 are a representative sum of the two spring rates. becomes. If the spring 98 characteristic is defined as curve A in FIG. 11 and the composite characteristic is defined as curve B, then the entire damper characteristic can be defined as curve C in FIG. This nonlinear damper spring rate has many desirable design and operating characteristics, and this is the desired result of the embodiment shown in FIG.

Referring now to FIG. 5, there is illustrated an embodiment utilizing the friction pad assembly of FIG. 4 and the dual inertia ring structure of FIG. In FIG. 5, reference numeral 90 again indicates the web of hub 92. In FIG. As also shown in FIG. 4, the cover plate 100 contacts a flat elastomeric ring 104 that cooperates with a flat friction material ring 106. As shown in FIG. Reference numeral 96' designates a first inertia ring of similar construction to the left half of inertia ring 96 of the embodiment of FIG. Reference numeral 112 designates a second inertia ring of similar construction to the right half of inertia ring 96 of the embodiment of FIG. Reference numeral 114 designates a disk of elastomeric material whose left side can be adhered to the right axial side of the inertial member 96'. Elastomer 11
The right side of 4 contacts a flat ring of friction material 116 of the same construction as that of friction material 106 described above. A labyrinth seal member 118 extends integrally from inertia ring 96' into a complementary recess in ring 112. Reference numeral 98 again indicates an elastomer band which is preferably also adhered to the inner radial circumference of the hub and inertia rings 96' and 112. The operating mode of the embodiment of FIG. 5 is as follows. The two inertia rings have different masses,
Each damps out a different specific frequency of torsional vibration within the operating range of the engine. Elastomeric disc-friction pad pairs 104-106 and 114-11 in the manner described above with respect to the sandwiched elastomeric and friction fabric members 104 and 106 of FIG.
6 are designed to frictionally damp vibrations and create mass interactions.

Referring now to FIG. 6, an embodiment similar to that of FIG. 5 is illustrated, except that three inertia rings are used. The embodiment of FIG. 6 is similar to the embodiment of FIG. 5, except for the additional inertia ring 120 sandwiched between inertia rings 96' and 112. Reference numeral 122 indicates an elastomeric disc in contact with elastomeric textile material 124. A similar elastomer disk 126 is attached to the inertia ring 1.
20 , the elastomer is also in contact with the friction material disc 128 . The mode of operation of the embodiment of FIG. 6 will be apparent from the description of the embodiment above.

Next, referring to FIG. 7, still another specific example is illustrated. Reference numeral 134 designates a web portion of a hub 136, which hub is adapted to be connected to a crankshaft having a rotation curve 14. Each axial surface of the hub has a contact element in the form of a flat ring member 138 attached by threaded fasteners 140 arranged at a plurality of angular intervals, two of which are shown. support. Reference numeral 144 designates an inertia ring whose two axial faces each have recesses 146 arranged with a plurality of angular spacings. The innermost portion of each recess supports a friction member in the form of a resilient spring 148 and a piston 150. This spring 148 connects the friction member 150 to the contact plate 138.
force into frictional contact with the innermost axial surface of the Reference numeral 152 designates an elastomeric band that is radially compressed by an inertial member, with both the radially innermost and radially outermost surfaces of the elastomer 152 being bonded to the respective ring and hub members. preferable.

The mode of operation of this embodiment will become clear. Simply put, the result of torsional vibration is the hub 136
〓〓〓〓
and the relative rotation between the inertia ring 144. This creates heat and results in elastomer 1
52 to attenuate torsional vibration. Additionally,
Such relative rotation causes friction piston member 1
50 and the contact plate 138. Such friction also serves to damp torsional vibrations. The embodiment of FIG. 7 is characterized by the use of dry friction, in which the time-averaged forces transmitted through the elastomer damping means in parallel configurations, ie, each means, are not equal. This type of damping action is again described in U.S. Pat.
It is of the type used by Krebs in No. 3075405. The improvement of the embodiment of FIG. 7 lies in the mechanical technique of its operation. A wipe or dry rub intermediate surface is seen between elements 150 and 138. Friction materials such as brake linings are low thermal conductors, so heat flow from frictional contact
Enter within 38. Due to its thin axial profile, it acts as a good heat conducting fin. It is generally known to utilize fins to cool hot objects as the larger surface area facilitates conduction to the surrounding air. Frictional contact at the axial ends of the inertia member 144 provides a mechanical lock against axial excursion of the inertia ring relative to the hub. Any additional external material that contaminates the friction surface, as well as worn material from the friction-related parts themselves, tends to be thrown out radially due to the increased centrifugal forces caused by the high rotational speed of the parts. .

Referring now to FIG. 8, an embodiment is illustrated that utilizes both elastomer damping and friction pad damping in the precise manner of FIG.
Reference numeral 160 designates a web portion of an annular hub 162, which hub 162 is adapted to be attached to a crankshaft having an axis of rotation 14. Reference numeral 164 designates a conical surface on the left hand portion of the outer rim of the hub, and reference numeral 166 designates a cylindrical portion connected to conical portion 166 and extending therefrom in the right axial plane of the torsional vibration damper. Reference numeral 168 indicates an elastomer band supported on the outer radial portion of the hub, which band supports the inertia ring 170. The right sides of both the inertia ring and hub are annularly rounded as shown. An elastomer band 168 is bonded to both the hub and the inertia ring and is radially deformed as shown by a bulge on the right side of the elastomer. Reference numeral 172 designates the radially outermost portion of the annular continuous clamping cup. The clamping cup exhibits a recess 173 disposed angularly around the plate, which is swaged radially inwardly into frictional contact with the illustrated outermost portion of the inertia ring 170. The clamping cup also includes a radially disposed portion 174, also angularly continuous, having an elastomer ring or disk on its inner surface illustrated by the reference numeral 176. To support. A ring or disc of friction material of the type described above has the reference numeral 17.
8. Both the elastomeric ring 176 and the friction material ring 178 fit at an angle within a continuous recess 180 on the left side of the hub 162. Again, since the friction coefficients of the intermediate surfaces are different, the elastomeric ring 17 is attached to the metal portion 174.
6 or the friction material 178 to the elastomer 176.

In operation, the device functions as described above with respect to the embodiment illustrated in FIG. Elastomeric band 1 whose inertia ring 170 damps vibrations in the manner described above.
68 and 162. At a relatively low torsional vibration inertia, the elastomer 176 and friction material 17
8 follows the vibration. However, as the torsional vibration rotational forces increase, the relatively small coefficient of friction between the material 178 and the metal of the hub 162 will result in relative motion between the two surfaces and between the rubber and the friction material. The relatively large coefficient of friction between the rubber and metal prevents relative rotation between these two parts. This result is the same as that produced in the example of FIG. 4 described above. However, separate fastening means are not required to position the elements in the configuration shown, simplifying assembly.

[Brief explanation of the drawing]

FIG. 1 is a partial sectional view of a multi-mass torsional vibration damper using a common elastic member and without interaction of inertial members (mass). FIG. 2 is a diagram similar to FIG. 1 of a specific example in which four inertial members are supported and there is no mass interaction. Figure 3 shows the dry friction mass interaction in a multi-mass vibration damper using a single elastic member.
FIG. 3 is a diagram of an example of a radially compressed rubber changing axially to produce a radially compressed rubber; FIG. 4 is a diagram of an embodiment having elastomer dry damping means in series in parallel with pure elastomer means to create a non-linear damper characteristic. FIG. 5 is a view similar to FIG. 4 in which two inertial members are used and the nonlinear mass interaction is caused by elastomeric dry friction damping means in series between the masses; FIG. 6 is a view similar to FIG. 5 in which the concept of FIG. 5 is applied to three inertial members. FIG. 7 is an illustration of an embodiment having a parallel combination of elastomeric and dry friction means with dry friction means additionally creating lateral damper cohesion. Figure 8 shows a shape that is relatively inexpensive to manufacture.
FIG. 4 is an illustration of an example of using various elastomeric cross-sectional shapes with a series of elastomeric dry friction damping media for the connectivity of dry friction damping means. Figures 9 and 10 are similar figures to Figure 1 with corresponding reference numbers and showing the case of no mass interaction; FIG. 11 is a diagram showing the linear and nonlinear behavior of some elements of FIG. 4, for example. 10...Torsional vibration damper, 12...Hub, 1
8... Elastomer band, 20... First inertial member, 24... Second inertial member, 41... Groove, 40
... groove, 50 ... hub, 52 ... elastomer band, 54, 56, 58 and 60 ... inertia ring,
62... assembly roundness, 64... groove, 70... hub, 80... elastomer band, 92... hub,
96...Inertia ring, 98...Elastomer band, 104...Annular elastomer. 〓〓〓〓

Claims (1)

[Claims] 1. Connected to the crankshaft of an internal combustion engine,
In a torsional vibration damper of the type having an annular hub 12, 50, 70 adapted to damp torsional vibrations of the crankshaft, a single elastomeric ring 1 surrounding the outer radial circumference of the hub;
8, 52, 80, and a plurality of annular inertia rings 2 each having an inner periphery engaged with and supported by the elastomeric ring on the hub.
0,24,54,56,58,60,20'2
4', wherein the inertia rings are frictionally connected to each other, and each of the inertia rings is angularly displaceable with respect to an adjacent inertia ring. 2. The torsional vibration damper of claim 1, wherein the elastomeric rings (18, 52) have a substantially uniform radial thickness over substantially all of their axial extent. 3. The torsional vibration damper of claim 1, wherein each inertia ring is axially spaced from its next adjacent inertia ring. 4 Connected to the crankshaft of the internal combustion engine,
In a torsional vibration damper of the type having an annular hub 92 adapted to damp torsional vibrations of the crankshaft and a single elastomeric ring 98 surrounding the outer radial circumference of the hub, the outer radial surface of the hub increases diametrically from a point on the surface intermediate the axial surface of the hub toward one axial surface of the hub, thereby forming a conical portion, each a plurality of annular inertia rings 96', 11 whose inner peripheries engage and are supported by the elastomeric rings on the hub;
2,120, the inertia rings being frictionally coupled to each other, each of the inertia rings being angularly displaceable with respect to an adjacent inertia ring, and one of the plurality of inertia rings 112 a flat elastomeric ring 114,1 having a conical inner radial periphery complementary in shape to and axially disposed along the conical portion around the hub;
22, 126 and substantially coextensive flat friction fabric rings 116, 124, 128 are sandwiched and compressed between the facing axial surfaces of the inertia rings. damper. 5 a plate member 100 attached to the hub, the plate member being in a radial plane and extending radially coextensive with the endmost axial surface of one of the inertia rings 96'; A flat elastic that has a portion forming a coextensive intermediate surface and is sandwiched and compressed between the coextensive intermediate surfaces in the radial direction.
5. The torsional vibration damper of claim 4 including a friction fabric ring 106 substantially coextensive with the toma ring 104. 6. The radially coextensive intermediate surfaces are located at the endmost axial surface of the hub opposite the endmost axial surface of the hub supporting the conical portion. The torsional vibration damper according to scope 5. 7. A claim in which the endmost inertial member with the coextensive portions comprises an annular recess in which the last-mentioned sandwiched elastomer and the friction member are at least partially located. The torsional vibration damper according to item 5. 8. The radially outermost circumference of at least one of the plurality of inertia rings includes an integral axially extending flange 118 that extends annularly into the next adjacent inertia ring 120, 112. 6. The torsional vibration damper of claim 5, wherein the damper is received by a groove thereby forming a labyrinth type seal for the sandwiched flat elastomer and the friction fabric member. 9. In a torsional vibration damper of the type comprising an annular hub 92 coupled to a crankshaft of an internal combustion engine for damping torsional vibrations of the crankshaft, at least one pair of inertia rings 96', 112; 96', 120; 120, 11
2 to the hub for limited relative rotation with respect to the hub and with respect to each other; means formed by a single elastomeric ring 98; Damping members 114 and 11 sandwiched and brought into contact
6,122,124,126,128,
the damping member is free to perform a back and forth oscillatory motion relative to the hub such that torque can be transferred from one of the paired inertia rings to the other of the paired inertia rings; A torsional vibration damper characterized by a curved shape. 10. The torsional vibration damper according to claim 9, wherein the damping member is a nonlinear elastic coupling. 11. Claim 10, wherein the non-linear coupling includes elastomeric discs 114, 122, 126 in contact with discs 116, 124, 128 of woven friction material, both discs being sandwiched between the inertia rings. Torsional vibration damper. 12 an annular hub 162 coupled to a crankshaft of an internal combustion engine and adapted to damp torsional vibrations of the crankshaft, the hub having an elastomeric band 16 on its outer radial surface;
In a torsional vibration damper of the type in which the elastomer band supports the outermost inertia ring 170, the outer radial surface of the hub 162 is ,
diametrically increasing toward one axial surface of the hub thereby forming a conical hub portion 164 and a complementary conical portion on the inner radial surface of the inertia ring 170; a continuous annular clamping attached to the outermost radial portion of the ring and extending radially inwardly across one axial face of the damper to form a radially disposed contact portion 174; Cup 172
and a flat friction fabric ring 178 adjacent and substantially coextensive with a flat elastomeric ring 176 sandwiched and compressed between the radially extending contact portion and the portion of the hub. A torsional vibration damper comprising: 13 The flat elastomer ring and the flat woven friction material ring form a continuous annular recess 18 in the hub.
Claim 12 located within 0
Torsional vibration damper as described in section. 14. The invention of claim 13, wherein the clamping cup is attached to the inertia ring by swaging the clamping cup at a plurality of annularly spaced portions. Torsional vibration damper. 15 having an annular hub 92 coupled to the crankshaft of the internal combustion engine and adapted to damp torsional vibrations of the crankshaft, the hub having an inertia to permit limited relative rotation with respect to the hub; In a torsional vibration damper of the type including means 98 for coupling rings 96, 96', a non-linear coupling 104, between said hub and said inertia ring.
106, the non-linear coupling being parallel to the means 98 for coupling the inertia ring to the hub;
A torsional vibration damper characterized in that the nonlinear coupling also connects the inertia ring to the hub. 16 The nonlinear coupling is a woven friction material circle〓〓〓〓
Claim 15 comprising an elastomer disc 104 in contact with a disc 106, both discs sandwiched between a member 100 secured to the hub 92 and the inertia rings 96, 96'. Torsional vibration damper as described. 17. The torsional vibration damper of claim 16, wherein said first-mentioned means for connecting said inertia ring to said hub is an elastomeric ring 98. 18 an annular hub 92, 162 connected to the crankshaft of the internal combustion engine and adapted to damp torsional vibrations of the crankshaft, the hub being adapted to allow limited relative rotation with respect to the hub; In a torsional vibration damper of the type including means for coupling an inertia ring 96, 96'170 to
Substantially coextensive disks of friction fabric material 106, 178
elastomeric disc 10 contacted by
4,176, the elastomeric disc and the friction material disc are secured to one of the hub and inertia ring by a member 100,172 and the hub. and an inertial member, thereby providing a non-linear coupling formed by the elastomeric disc and the friction material disc between the hub and the inertial ring. vibration damper.