Patent Publication Number: US-2022228494-A1

Title: Assembly for a turbomachine

Description:
FIELD OF THE INVENTION 
     The present invention relates to an assembly for a turbomachine. 
     The invention relates more specifically to an assembly for a turbomachine comprising a damper. 
     STATE OF THE ART 
     A turbomachine known from the state of the art comprises a casing and a fan capable of being rotated relative to the casing, around a longitudinal axis, by means of a fan shaft. The fan comprises a disk centered on the longitudinal axis, and a plurality of blades distributed circumferentially at the outer part of the disk. 
     The range of operation of the fan is limited. More specifically, the evolution of a compression rate of the fan as a function of an air flow rate it draws when rotated, is restricted to a predetermined range. 
     Beyond this range, the fan is indeed subjected to aeroelastic phenomena which destabilize it. More specifically, the air circulating through the running fan supplies energy to the blades, and the blades respond in their eigenmodes at levels that may exceed the endurance limit of the material constituting them. This fluid-structure coupling therefore generates vibrational instabilities which accelerate the wear of the fan and reduce its service life. 
     A fan which comprises a reduced number of blades, and which is subjected to high aerodynamic loads, is very sensitive to this type of phenomena. 
     This is the reason why it is necessary to guarantee a sufficient margin between the stable operating range and the areas of instability, so as to spare the endurance limits of the fan. To do so, it is known practice to equip the fan with dampers. Examples of dampers have been described in documents FR 2 949 142, EP 1 985 810 and FR 2 923 557, in the name of the Applicant. These dampers are all configured to be housed between the platform and the root of each blade, within the housing delimited by the respective stilts of two successive blades. Furthermore, such dampers operate during a relative movement between two successive blade platforms, by dissipation of the vibration energy, for example by friction. Consequently, these dampers focus only on damping a first vibratory mode of the blades which characterizes a synchronous response of the blades to the aerodynamic loads. In this first vibratory mode, the inter-blade phase-shift is non-zero. 
     However, such dampers are totally ineffective for damping a second vibratory mode in which each blade flaps relative to the disk with a zero inter-blade phase-shift. Indeed, in this second vibratory mode, there is no relative movement between two successive blade platforms. This particular response of the blades to the aerodynamic loads, although asynchronous, still involves a non-zero moment on the fan shaft. In addition, this second vibratory mode is coupled between the blades, the disk and the fan shaft. The amplitude of this second vibratory mode is all the more important as the blades are large. 
     There is therefore a need to overcome at least one of the drawbacks of the state of the art described above. 
     DISCLOSURE OF THE INVENTION 
     One aim of the invention is to damp a mode of vibration of a rotor in which the phase-shift between the blades of said rotor is zero. Another aim of the invention is to influence the damping of modes of vibration of a rotor in which the phase-shift between the blades of said rotor is non-zero. Another aim of the invention is to propose a damping solution which is simple and easy to implement. 
     To this end, according to a first aspect of the invention, an assembly for a turbomachine is proposed, comprising: 
     a casing, 
     a first rotor: 
     movable in rotation relative to the casing around a longitudinal axis, and 
     comprising: 
     a disk, and 
     a plurality of blades capable of flapping relative to the disk during a rotation of the first rotor relative to the casing, 
     a second rotor movable in rotation relative to the casing around the longitudinal axis, and 
     a damper configured to damp a movement of the first rotor relative to the second rotor, in a plane orthogonal to the longitudinal axis, the movement being caused by a flapping of at least one blade among the plurality of blades, the damper comprising: 
     a first part bearing on the first rotor, and having: 
     a first radially inner surface extending all around the longitudinal axis, 
     a first radially outer surface extending all around the first radially inner surface, and 
     a first radial thickness measured perpendicular to the longitudinal axis between the first radially inner surface and the first radially outer surface, 
     a second part bearing on the second rotor, and having: 
     a second radially inner surface extending all around the longitudinal axis, 
     a second radially outer surface extending all around the second radially inner surface, and 
     a second radial thickness measured perpendicular to the longitudinal axis between the second radially inner surface and the second radially outer surface, and 
     a third part connecting the first part to the second part, and having: 
     a third radially inner surface extending all around the longitudinal axis, 
     a third radially outer surface extending all around the third radially inner surface, and 
     a third radial thickness measured perpendicular to the longitudinal axis between the third radially inner surface and the third radially outer surface, 
     in which the third radial thickness is greater than at least one among the first radial thickness and the second radial thickness and the third part comprises a bulge. 
     It is by damping a movement of the first rotor relative to the second rotor, in a plane orthogonal to the longitudinal axis, that it is possible to influence the second vibratory mode. Actually, unlike the first vibratory mode, the second vibratory mode is characterized by a zero inter-blade phase-shift. Consequently, placing a damper between two successive blades of a rotor, as it has already been proposed in the prior art, has no effect on the second vibratory mode. The damper of the assembly described above has, for its part, the advantage of influencing the second vibratory mode because it plays on an effect of the second vibratory mode: the movement of the first rotor relative to the second rotor, in the plane orthogonal to the longitudinal axis. By opposing this effect, the damper disrupts the cause thereof that is to say dampens the second vibratory mode. It should nevertheless be noted that the first vibratory mode also participates in the movement of the first rotor relative to the second rotor, in the plane orthogonal to the longitudinal axis. Consequently, by opposing this effect, the damper also participates in disrupting another cause thereof that is say damping the first vibratory mode. In addition, since the damper is annular, it allows distributing the bearing stresses applied by the damper on the first rotor and on the second rotor, over a larger surface. From there, the damper wears less the first rotor and the second rotor on which it bears. Finally, as the third part is thicker than the first part and the second part, it is more massive. The third part therefore allows limiting the tangential propagation of the vibratory modes to which the first rotor and the second rotor are subjected. Thus, the damper is capable, thanks to this third part, of dissipating the vibrations by its work in bending and in inertia. 
     Advantageously, but optionally, the assembly according to the invention may further comprise one of the following characteristics, taken alone or in combination with one or several of the other of the following characteristics: 
     in such an assembly: 
     the first part is configured to apply a first centrifugal force on the first rotor, and 
     the second part is configured to apply a second centrifugal force on the second rotor, 
     the first bearing part has a radially outer surface coming into contact with a radially inner surface of the first rotor and the second bearing part has a radially outer surface coming into contact with a radially inner surface of the second rotor, 
     the third radial thickness is greater than each among the first radial thickness and of the second radial thickness, 
     the second radial thickness is greater than the first radial thickness, 
     the bulge comprises a first lip protruding radially inwardly from the damper, 
     the bulge comprises a second lip protruding radially outwardly from the damper, 
     the third part comprises a depression, 
     in such an assembly: 
     the third part has a first bearing surface arranged to apply a first force on the second rotor, the first force having a first longitudinal component in a first direction parallel to the longitudinal axis, and a first radial component in a second direction orthogonal to the longitudinal axis, the first longitudinal component being greater than the first radial component, 
     the second part has a second bearing surface arranged to apply a second force on the second rotor, the second force having a second longitudinal component in the first direction, and a second radial component in the second direction, the second radial component being greater than the second longitudinal component, 
     each of the blades among the plurality of blades comprises: 
     a blade root connecting the blade to the disk, 
     a profiled blading, 
     a stilt connecting the blading to the blade root, and 
     a platform connecting the blading to the stilt and extending transversely to the stilt, the first bearing part bearing on each of the platforms of the blades among the plurality of blades, 
     the second rotor comprises a shroud, the shroud comprising a circumferential extension, the second bearing part bearing on the circumferential extension, and 
     the damper is annular, and extends all around the longitudinal axis. 
     According to a second aspect of the invention, there is proposed a turbomachine comprising an assembly as described above, and in which the first rotor is a fan and the second rotor is a low-pressure compressor. 
    
    
     
       DESCRIPTION OF THE FIGURES 
       Other characteristics, aims and advantages of the invention will emerge from the following description, which is purely illustrative and not limiting, and which should be read in relation to the appended drawings in which: 
         FIG. 1  schematically illustrates a turbomachine, 
         FIG. 2  comprises a sectional view of a part of a turbomachine, and a curve indicating a tangential movement of different elements of this turbomachine part as a function of the position of said elements along a longitudinal axis of the turbomachine, 
         FIG. 3  is a sectional view of part of an exemplary embodiment of an assembly according to the invention, 
         FIG. 4  is a perspective view of part of an exemplary embodiment of an assembly according to the invention, and 
         FIG. 5  is a perspective view of a part of a damper of an exemplary embodiment of an assembly according to the invention. 
     
    
    
     In all of the figures, the similar elements bear identical references. 
     DETAILED DESCRIPTION OF THE INVENTION 
     Turbomachine  1   
     Referring to  FIG. 1 , a turbomachine  1  comprises a casing  10 , a fan  12 , a low-pressure compressor  140 , a high-pressure compressor  142 , a combustion chamber  16 , a high-pressure turbine  180  and a low-pressure turbine  182 . 
     Each of the fan  12 , of the low-pressure compressor  140 , of the high-pressure compressor  142 , of the high-pressure turbine  180  and of the low-pressure turbine  182  is movable in rotation relative to the casing  10  around a longitudinal axis X-X. 
     In the embodiment illustrated in  FIG. 1 , and as also visible in  FIGS. 2 and 3 , the fan  12  and the low-pressure compressor  140  are secured in rotation and are capable of being rotated by a low-pressure shaft  13  which is itself capable of being rotated by the low-pressure turbine  182 . The high-pressure compressor  142  is for its part capable of being rotated by a high-pressure shaft  15 , which is itself capable of being rotated by the high-pressure turbine  180 . 
     In operation, the fan  12  draws in an air stream  110  which separates between a secondary stream  112  circulating around the casing  10 , and a primary stream  111  successively compressed within the low-pressure compressor  140  and the high-pressure compressor  142 , ignited within the combustion chamber  16 , then successively expanded within the high-pressure turbine  180  and the low-pressure turbine  182 . 
     The upstream and the downstream are here defined relative to the direction of normal air flow  110 ,  111 ,  112  through the turbomachine  1 . Likewise, an axial direction corresponds to the direction of the longitudinal axis X-X, a radial direction is a direction which is perpendicular to this longitudinal axis X-X and which passes through said longitudinal axis X-X, and a circumferential or tangential direction corresponds to the direction of a planar and closed curved line, all the points of which are at equal distance from the longitudinal axis X-X. Finally, and unless otherwise specified, the terms “inner (or internal)” and “outer (or external)”, respectively, are used with reference to a radial direction such that the inner (i.e. radially inner) part or face of an element is closer to the longitudinal axis X-X than the outer (i.e. radially outer) part or face of the same element. 
     Fan  12  and Low-Pressure Compressor  140   
     Referring to  FIGS. 1 to 3 , the fan  12  comprises a disk  120  and a plurality of blades  122  circumferentially distributed at an outer part of the disk  120 . 
     Referring to  FIGS. 2 and 3 , each of the blades  122  of the plurality of blades  122  comprises: 
     a blade root  1220  connecting the blade  122  to the disk  120 , 
     a profiled blading  1222 , 
     a stilt  1224  connecting the blading  1222  to the blade root  1220 , and 
     a platform  1226  connecting the blading  1222  to the stilt  1224  and extending transversely to the stilt  1224 . 
     The blade root  1220  may be integral with the disk  120  when the fan  12  is a one-piece bladed disk. Alternatively, as seen in  FIG. 3 , the blade root  1220  can be configured to be housed in a cell  1200  of the disk  120  provided for this purpose. 
     As seen in  FIGS. 2 and 3 , the low-pressure compressor  140  also comprises a plurality of blades  1400  fixedly mounted at an outer part of a shroud  1402 , said shroud  1402  comprising a circumferential extension  1404  at the outer end from which radial sealing wipers  1406  extend. The radial sealing wipers  1406  face the platforms  1226  of the blades  122  of the fan  12 , so as to guarantee the inner sealing of the flowpath within which the primary stream  111  circulates. As more specifically visible in  FIG. 3 , the shroud  1402  of the low-pressure compressor  140  is fixed to the disk  120  of the fan  12 , for example by bolting. 
     Each of the blades  122  of the plurality of fan  12  blades  122  is capable of flapping, by vibrating relative to the disk  120  during a rotation of the fan  12  relative to the casing  10 . More specifically, during the coupling between the air  110  circulating within the fan  12  and the profiled bladings  1222 , the blades  122  are the site of aeroelastic floating phenomena on different vibratory modes, and whose amplitude may be such that it exceeds the endurance limits of the materials constituting the fan  12 . These vibratory modes are furthermore coupled to the opposite compressive forces upstream of the turbomachine  1 , and to the expansion forces downstream of it. 
     A first vibratory mode characterizes a synchronous response of the blades  122  to the aerodynamic loads, in which the inter-blade phase-shift is non-zero. 
     A second vibratory mode characterizes an asynchronous response of the blades  122  to the aerodynamic loads, in which the inter-blade phase-shift is zero. The amplitude of the flapping of the second vibratory mode is moreover as large as the fan  12  blades  122  are large. Furthermore, this second vibratory mode is coupled between the blades  122 , the disk  120  and the fan shaft  13 . The frequency of the second vibratory mode is in addition one and a half times greater than that of the first vibratory mode. Finally, the second vibratory mode has a nodal deformation at mid-height of the fan  12  blades  122 . 
     In vibratory modes, including the second vibratory mode, the flapping of the blades  122  involves a non-zero moment on the low-pressure shaft  13 . In particular, these vibratory modes cause intense torsional forces within the low-pressure shaft  13 . 
     The vibrations induced by the flapping of the blades  122  of the fan  12 , but also by the flapping of the blades  1400  of the low-pressure compressor  140 , lead to significant relative tangential movements between the fan  12  and the low-pressure compressor  140 . Indeed, the length of the blades  122  of the fan  12  is greater than the length of the blades  1400  of the low-pressure compressor  140 . Consequently, the tangential bending moment caused by the flapping of a blade  122  of the fan  12  is greater than the tangential bending moment caused by flapping of a blade  1400  of the low-pressure compressor  140 . The blading of the blades  122  of the fan  12  and of the blades  1400  of the low-pressure compressor  140  then have very different behaviors. Furthermore, the mounting stiffness within the fan  12  is different from the mounting stiffness within the low-pressure compressor  140 . 
     As seen more specifically in  FIG. 2 , this results in particular in a large-amplitude movement of the fan  12  relative to the low-pressure compressor  140 , in a plane orthogonal to the longitudinal axis X-X, at the interface between the platforms  1226  of the blades  122  of the fan  12  and the radial sealing wipers  1406  of the circumferential extension  1404  of the shroud  1402  of the low-pressure compressor  140 . The amplitude of this movement for the second vibratory mode is for example between 0.01 and 0.09 millimeter, typically on the order of 0.06 millimeter, or, in another example, on the order of a few tenths of a millimeter, for example 0.1 or 0.2 or 0.3 millimeter. 
     Damper  2   
     A damper  2  is used to damp these vibrations of the fan  12  and/or of the low-pressure compressor  140 . 
     The damper  2  is in particular configured to damp a movement of the fan  12  relative to the low-pressure compressor  140 , in a plane orthogonal to the longitudinal axis X-X, the movement being caused by a flapping of at least one blade  122  among the plurality of blades  122  of the fan  12 . 
     Referring to  FIGS. 3 to 5 , the damper  2  comprises: 
     a first part  21  bearing on the fan  12 , 
     a second part  22  bearing on the low-pressure compressor  140 , and 
     a third part  23  connecting the first part  21  to the second part  22 . 
     As in particular seen in  FIG. 5 , the damper  2  is annular, and therefore extends all around the longitudinal axis X-X. More specifically, the first part  21  has a first radially inner surface  211  extending all around the longitudinal axis X-X, and a first radially outer surface  212  extending all around the first radially inner surface  211 . In addition, the second part  22  has a second radially inner surface  221  extending all around the longitudinal axis X-X, and a second radially outer surface  222  extending all around the second radially inner surface  221 . Finally, the third part  23  has a third radially inner surface  2310  extending all around the longitudinal axis X-X, and a third radially outer surface  2320  extending all around the third radially inner surface  2310 . 
     In addition, as seen in  FIG. 4 , the first part  21  has a first radial thickness E 1  measured perpendicular to the longitudinal axis X-X between the first radially inner surface  211  and the first radially outer surface  212 . Likewise, the second part  22  has a second radial thickness E 2  measured perpendicular to the longitudinal axis X-X between the second radially inner surface  221  and the second radially outer surface  222 . Finally, the third part  23  has a third radial thickness E 3  measured perpendicular to the longitudinal axis X-X between the third radially inner surface  2310  and the third radially outer surface  2320 . 
     The third radial thickness E 3  is greater than at least one of the first radial thickness E 1  and of the second radial thickness E 2 . In one embodiment, for example illustrated in  FIG. 4 , the third radial thickness E 3  is greater than each of the first radial thickness E 1  and of the second radial thickness E 2 . In this way, the third part  23  is more massive than the first part  21  and than the second part  22 . In an also advantageous variant, the second radial thickness E 2  is greater than the first radial thickness E 1 , so as to promote the bearing of the second part  22  on the low-pressure compressor  140 . 
     In one advantageous embodiment, the first part  21  bears on each of the platforms  1226  of the blades  122  of the fan  12 , preferably at an inner surface of each of the platforms  1226 . An annular damper  2  is moreover particularly suitable for a fan  12  comprising a disk  120  which is integrally bladed. Indeed, in a fan  12  where the blades  122  are added onto the disk  120 , if the damper  2  is annular, then the bearing of the first part  21  on the different platforms  1226  of the blades  122  is not uniform. This induces inhomogeneous damping around the longitudinal axis X-X and, hence, risks of wear of the platforms  1226  and of the damper  2 . The inner surfaces of the platforms  1226  may include reliefs so as to be axisymmetric. This circumferential non-symmetry on the internal side of the platforms  1226  can thus optimize the mutual bearings of the damper  2 , particularly their distributions, while favoring, where appropriate, bearing wears on these reliefs. 
     In addition, the second part  22  bears on the circumferential extension  1404  of the shroud  1402  of the low-pressure compressor  140 , at an inner surface of the radial sealing wipers  1406 . Indeed, it is in this position that the movement of the fan  12  relative to the low-pressure compressor  140 , in the plane orthogonal to the longitudinal axis X-X, is of greater amplitude, typically a few millimeters. Consequently, the damper  2  is particularly effective there. 
     In one embodiment, the damper  2  comprises a material from the range having the trade name “SMACTANE® ST” and/or “SMACTANE® SP”, for example a material of the type “SMACTANE® ST 70” and/or “SMACTANE® SP 50”. It has indeed been observed that such materials have suitable damping properties. 
     Referring to  FIG. 3 , in one embodiment, the first part  21  is configured to apply a first centrifugal force Cl on the fan  12 , while the second part  22  is configured to apply a second centrifugal force C 2  on the low-pressure compressor  140 . To apply the first centrifugal force C 1 , the first bearing part  21  has a radially outer surface coming into contact with a radially inner surface of the fan  12 , typically a radially inner surface of the platform  1226 . In order to apply the second centrifugal force C 2 , the second bearing part  22  has a radially outer surface coming into contact with a radially inner surface of the low-pressure compressor  140 , typically a radially inner surface of the circumferential extension  1404 , for example a radially inner surface of the sealing wipers  1406 . In this way, these parts  21 ,  22  are each dynamically coupled respectively to the fan  12  and to the low-pressure compressor  140  on which each bears, so as to undergo the same vibrations as each of the fan  12  and of the low-pressure compressor  140 . 
     The third part  23  is stiffer, in particular in a tangential direction. Thus, in operation, a movement of the fan  12  relative to the low-pressure compressor  140 , in a plane orthogonal to the longitudinal axis X-X, causes a tangential shear of the damper  2  which leads to circumferential movements of said damper  2 . The respective bearings on the fan  12  and the low-pressure compressor  140  are therefore interrupted, then quickly resumed to apply again the centrifugal forces C 1 , C 2 . These interruptions and resumptions of the bearings allow the damping. Advantageously, the tangential movements of the high-frequency fan  12  are damped when the parts  21 ,  22  are bearing against the fan  12  and the low-pressure compressor  140 . The interruption of the bearings, then the circumferential sliding, allows damping lower frequencies. In this way, the damper  2  is effective over a wide range of frequencies. 
     Referring to  FIG. 4 , in one embodiment, the third part  23  comprises a preferably annular bulge  231 ,  232 . Advantageously, the bulge  231 ,  232  comprises a first lip  231 , itself also annular, and radially protruding inwardly from the damper  2 . The first lip  231  is intended to make the third part  23  heavier, which advantageously increases its tangential inertia. 
     Alternatively or additionally as illustrated in  FIG. 4 , the bulge  231 ,  232  comprises a second lip  232 , also annular, and radially protruding outwardly from the damper  2 . In addition to its function of weighing down the third part  23  which advantageously leads to an increase in the tangential rigidity, the second lip also allows ensuring the axial setting of the damper  2  between the fan  12  and the low-pressure compressor  140 . 
     Referring to  FIG. 4 , in one embodiment: 
     the third part  23  has a first bearing surface  2321  arranged to apply a first force F 1  on the low-pressure compressor  140 , the first force F 1  having a first longitudinal component F 1 L in a first direction parallel to the longitudinal axis X-X, and a first radial component F 1 R in a second direction orthogonal to the longitudinal axis X-X, the first longitudinal component F 1 L being greater than the first radial component F 1 R, 
     the second part  22  has a second bearing surface  2200  arranged to apply a second force F 2  on the low-pressure compressor  140 , the second force F 2  having a second longitudinal component F 2 L in the first direction, and a second radial component F 2 R in the second direction, the second radial component F 2 R being greater than the second longitudinal component F 2 L. 
     In other words, the third part  23  ensures the axially positioned bearing of the damper  2 , via the first bearing surface  2321 , since it is a downstream axial surface of the damper  2  coming into contact with an upstream axial surface of the low-pressure compressor  140 . 
     Furthermore, the second part  22  ensures the radially positioned bearing of the damper  2 , via the second bearing surface  2200 , since it is a radially outer surface of the damper  2  coming into contact with a radially inner surface of the low-pressure compressor  140 . In addition, in operation, the second bearing surface  2200  participates in the application of the second centrifugal force C 2  on the low-pressure compressor  140 . Advantageously, it is the second lip  232  of the third part  23  which has the first bearing surface  2321 , as seen in  FIG. 4 . Referring to  FIGS. 4 and 5 , in one embodiment, the third part  23  comprises a depression  233 , preferably an annular depression. The depression  233  can be made at an outer surface  2320  or an inner surface  2310  of the third part  23 , upstream or downstream of the bulge  231 ,  232 . In the embodiment illustrated in  FIG. 5 , the depression  233  extends upstream of the bulge. When the depression  233  extends downstream of the bulge  231 ,  232 , as illustrated in  FIG. 4 , at an outer surface  2320  of the third part  23 , it ensures a clearance which allows the damper  2  to avoid to rub on one corner of the radial sealing wipers  1406 . In any event, the depression  233  promotes the axial setting of the damper  2  between the fan  12  and the low-pressure compressor  140 , but also the sealing of the flowpath of the primary air stream  111 . Indeed, under the effect of the first centrifugal force C 1 , the first part  21  can thus be compressed downstream. 
     In one embodiment, one at least of the first part  21 , the second part  22  and the third part  23  comprises an additional coating configured to reduce the friction and/or the wear of the fan and/or of the low-pressure compressor  140 . This additional coating is fixedly mounted on an outer surface of the damper  2 , for example by bonding. The additional coating is of the dissipative and/or viscoelastic and/or damping type. It may indeed comprise a material from the range having the trade name “SMACTANE® ST” and/or “SMACTANE® SP”, for example a material of the type “SMACTANE® ST 70” and/or “SMACTANE® SP 50”. It can also comprise a material chosen from those having mechanical properties similar to those of Vespel, Teflon or any other material with lubricating properties. More generally, the additional coating material advantageously has a coefficient of friction between 0.3 and 0.07. The coating allows in particular increasing the tangential stiffness of the damper  2  when, in operation, it applies the centrifugal forces C 1 , C 2  so that the movement of the fan  12  relative to the low-pressure compressor  140 , in the plane orthogonal to the longitudinal axis X-X, is damped by energy dissipation by means of a viscoelastic shear of its coating. 
     In one embodiment, one at least of the first part  21 , the second part  22  and the third part  23  is treated by dry lubrication, with a view to maintaining the value of the coefficient of friction between the damper  2  and either or both of the fan  12  and of the low-pressure compressor  140 . This material with lubricating properties is for example of the MoS 2  type. 
     In all that has been described above, the damper  2  is configured to damp a movement of the fan  12  relative to the low-pressure compressor  140 , in the plane orthogonal to the longitudinal axis X-X. 
     This is however not limiting, since the damper  2  is also configured to damp a movement of any first rotor  12  relative to any second rotor  140 , in a plane orthogonal to the longitudinal axis X-X, as long as the first rotor  12  is movable in rotation relative to the casing  10  around the longitudinal axis X-X and comprises a disk  120  as well as a plurality of blades  122  capable of flapping by vibrating relative to the disk  120  during a rotation of the first rotor  12  relative to the casing  10 , and as the second rotor  140  is also movable in rotation relative to the casing  10  around the longitudinal axis X-X. 
     Thus, the first rotor  12  can be a first stage of the high-pressure compressor  142  or of the low-pressure compressor  140 , and the second rotor  140  can be a second stage of said compressor  140 ,  142 , successive to the first stage of compressor  140 ,  142 , upstream or downstream thereof. Alternatively, the first rotor  12  can be a first stage of a high-pressure turbine  180  or of low-pressure turbine  182 , and the second rotor  140  can be a second stage of said turbine  180 ,  182 , successive to the first stage of turbine  180 ,  182 , upstream or downstream thereof. 
     In any event, the damper  2  has a small space requirement. Consequently, it can be easily integrated into the existing turbomachines. 
     In addition, by being configured to exert centrifugal forces C 1 , C 2  on the first rotor  12  and on the second rotor  140 , the damper  2  ensures significant tangential stiffness between the first rotor  12  and the second rotor  140 . It thus differs from an excessively flexible damper which would only deform during a movement of the first rotor  12  relative to the second rotor  140 , in the plane orthogonal to the longitudinal axis X-X. On the contrary, the damper  2  dissipates such a movement: 
     either by friction and/or oscillations between a state where the damper  2  is bonded on the rotors  12 ,  140  and a state where the damper  2  slides on the rotors  12 ,  140 , which allows damping in particular the low frequencies, 
     or by viscoelastic shear within the damper  2 , which allows damping in particular the high frequencies. 
     However, the damper  2  remains flexible enough to maximize the contact surfaces between said damper  2  and the rotors  12 ,  140  on which it bears. To do so, the damper  2  has a tangential rigidity greater than an axial rigidity and a radial rigidity. 
     The contact forces between the damper  2  and the rotors  12 ,  140  can in particular be adjusted by means of additional coatings. At low frequencies, it is indeed necessary to ensure that the centrifugal forces C 1 , C 2  exerted by the damper  2  on the rotors  12 ,  140  are not too large, in order to guarantee that the damper  2  can oscillate between a bonded state and a slippery state on the rotors  12 ,  140 , and thus damp by friction. At high frequencies, on the other hand, it is necessary to ensure that the centrifugal forces C 1 , C 2  exerted by the damper  2  on the rotors  12 ,  140  are sufficiently large for the pre-stress of the damper  2  on the rotors  12 ,  140  to be sufficient, in order to ensure that the damper  2  can be the viscoelastic shear seat. 
     The wear of the rotors  12 ,  140  is in particular limited by the treatment of the surfaces of the damper  2  bearing on the rotors  12 ,  140 , for example to equip them with a coating with a low coefficient of friction.