Patent Publication Number: US-10309224-B2

Title: Split ring spring dampers for gas turbine rotor assemblies

Description:
CROSS-REFERENCE TO RELATED APPLICATION 
     This application claims the benefit of priority under 35 U.S.C. § 119(e) to U.S. Provisional Application No. 62/004,362, filed May 29, 2014, which is incorporated herein by reference in its entirety. 
    
    
     BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     The present disclosure relates to vibration damping, and more particularly to mechanical damping devices for gas turbine engine components. 
     2. Description of Related Art 
     Gas turbine engines ignite compressed air and fuel to create a flow of hot combustion gases that drive multiple stages of turbine blades. The turbine blades extract energy from the flow of hot combustion gases to drive a turbine rotor. The turbine rotor drives a fan to provide thrust and a compressor to provide a flow of compressed air. Disk covers coupled to the turbine blade stages form an inner portion of a gas path traversed by the hot combustion gases. These covers provide separation between the hot combustion gases traversing the turbine disk and portions of the disk not exposed to the combustion gases. 
     Turbine stage disk covers can be subject to vibrational forces and/or flutter due to fluid flow pulsation during engine operation. These forces can require damping, typically through cover geometry and/or material selection, or through use of a mechanical damper. Mechanical dampers function by absorbing vibrational energy through mechanical contact with the damped structure to reduce the response of the damped structure from vibrational forces and/or flutter otherwise resulting from fluid flow passed the structure during engine operation. 
     Such conventional methods and systems have generally been considered satisfactory for their intended purpose. However, there is still a need in the art for improved mechanical damper. There is also a need for improved dampers with increased ability to withstand engine transportation loads. The present disclosure provides a solution for this need. 
     SUMMARY OF THE INVENTION 
     A spring damper includes a split ring body. The split ring body defines a center and a circular gap separating opposed first and second end portions of the split ring body. The first and second end portions are connected by an evenly spaced segment of the split ring body that is evenly spaced from the body center. At least one of the first and second end portions is unevenly spaced from the center in relation to the evenly spaced segment. 
     In certain embodiments, the evenly spaced segment can be offset from the center by a uniform radius. An end of the first end portion can be spaced radially outward from the center in relation to the evenly spaced segment. An end of the second end portion can be spaced radially outward from the center in relation to the evenly spaced segment. It is contemplated both ends of the end portions can be spaced radially outward from the center in relation to the evenly spaced segment. 
     In accordance with certain embodiments the split ring body can have an arcuate shape, such as a circular or elliptical shape for example. The evenly spaced segment can span an arc extending about 270 degrees around the center of the split ring body. At least one of the first and second end portions can transition to a larger radius of curvature relative to the evenly spaced segment within a span of about 0 degrees to 180 degrees of the split ring body. 
     It is also contemplated that in certain embodiments the spring damper can have an unloaded configuration wherein the end portion ends extend radially outward in relation to the evenly spaced segment and define an unloaded gap width therebetween. The spring damper can also have a statically loaded configuration wherein the end portion ends are spaced radially inward in relation to the evenly spaced segment and define a statically loaded gap width therebetween. The statically loaded gap width can be less than the unloaded gap width. 
     It is further contemplated that the spring damper can have a dynamically loaded configuration wherein end portion ends and the evenly spaced segment are equidistantly spaced about the center. End portion ends can be separated by a gap with a dynamically loaded gap width therebetween that is greater than the statically loaded gap width. The dynamically loaded gap width can also be less than the unloaded gap width. 
     A rotor stage includes a disk, a disk cover and a spring damper as described above. The disk cover is connected to the disk and the spring damper is connected to the disk cover. The disk cover imparts a preload into the split ring body by exerting preload forces on the first and second end portions of the spring damper such that the first and second end portions are spaced radially inward toward the center by at least the same distance as the evenly spaced segment. In accordance with certain embodiments the preload forces can be such that ends of the end portions are spaced radially inward toward to the center in relation to the evenly spaced segment. 
     These and other features of the systems and methods of the subject disclosure will become more readily apparent to those skilled in the art from the following detailed description of the preferred embodiments taken in conjunction with the drawings. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       So that those skilled in the art to which the subject disclosure appertains will readily understand how to make and use the devices and methods of the subject disclosure without undue experimentation, preferred embodiments thereof will be described in detail herein below with reference to certain figures, wherein: 
         FIG. 1  is a schematic, partial cross-sectional side view of an exemplary embodiment of a gas turbine engine constructed in accordance with the present disclosure, showing a rotor stage; 
         FIG. 2  is a schematic, cross-sectional side view of a portion of the gas turbine engine of  FIG. 1 , showing the rotor stage and a disk, a disk cover, and a spring damper of the rotor stage; 
         FIG. 3  is a schematic axial view of the spring damper of  FIG. 2 , showing an evenly spaced segment and end portions of the spring damper; 
         FIG. 4  is a schematic axial view of the spring damper of  FIG. 3 , showing the spring damper in an unloaded configuration; 
         FIG. 5  is a schematic axial view of the spring damper of  FIG. 3 , showing the spring damper in a statically loaded configuration; and 
         FIG. 6  is a schematic axial view of the spring damper of  FIG. 3 , showing the damper in a dynamically loaded configuration. 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Reference will now be made to the drawings wherein like reference numerals identify similar structural features or aspects of the subject disclosure. For purposes of explanation and illustration, and not limitation, a partial view of an exemplary embodiment of a gas turbine engine including the spring damper in accordance with the disclosure is shown in  FIG. 1  and is designated generally by reference character  10 . Other embodiments of gas turbine engines and spring dampers for gas turbine engines in accordance with the disclosure, or aspects thereof, are provided in  FIGS. 2-6 , as will be described. Embodiments of spring dampers described herein can be used for damping components in aircraft gas turbine engines, terrestrial gas turbines, and marine gas turbines. 
     As used herein, the term dynamically loaded refers loading imposed on engine components when engine rotary components are rotating during engine operation. Transportation load refers to loads exerted on engine rotary components when the rotary components are not rotating. This includes time intervals during which the engine is not operating, such as when the engine or engine subassembly is being transported as a spare for example. 
       FIG. 1  schematically illustrates gas turbine engine  10 . Gas turbine engine  10  as disclosed herein as a two-spool turbofan that generally incorporates a fan section  22 , a compressor section  24 , a combustor section  26  and a turbine section  28 . Fan section  22  drives air along a bypass flow path B in a bypass duct defined within a nacelle  15 , while the compressor section  24  drives air along a core flow path C for compression and communication into combustor section  26  followed by expansion through turbine section  28 . Although depicted as a two-spool turbofan gas turbine engine in the disclosed non-limiting embodiment, it should be understood that the concepts described herein are not limited to use with two-spool turbofans as the teachings may be applied to other types of turbofan engines including three-spool engine architectures. 
     Exemplary gas turbine engine  10  generally includes a low-speed spool  30  and high-speed spool  32  mounted for rotation about an engine rotational axis A relative to an engine static structure  36  via several bearing systems  38 . It should be understood that various bearing systems  38  at various locations may alternatively or additionally be provided, and the location bearing systems  38  may be varied as appropriate to the application. 
     Low-speed spool  30  generally includes an inner shaft  40  that interconnects a fan  42 , a first (or low-pressure) compressor  44  and a first (or low-pressure) turbine  46 . Inner shaft  40  is connected to fan  42  through a speed change mechanism, which in exemplary gas turbine engine  10  is illustrated as a geared architecture  48  to drive fan  42  at a lower speed than low-speed spool  30 . High-speed spool  32  includes an outer shaft  50  that interconnects a second (or high-pressure) compressor  52  and a second (or high-pressure) turbine  54 . A combustor  56  is arranged in exemplary gas turbine engine  10  between high-pressure compressor  52  and high-pressure turbine  54 . A mid-turbine frame  57  of engine static structure  36  is arranged generally between high-pressure turbine  54  and low-pressure turbine  46 . Mid-turbine frame  57  further supports bearing systems  38  in turbine section  28 . Inner shaft  40  and outer shaft  50  are concentric and rotate via bearing systems  38  about engine central rotation axis A which is collinear with their rotation axes. 
     Core airflow is compressed by low-pressure compressor  44 , further compressed by high-pressure compressor  52 , mixed and burned with fuel in combustor  56 , and expanded over high-pressure turbine  54  and low-pressure turbine  46 . Mid-turbine frame  57  includes airfoils  59 , which are in core airflow path C. Low-pressure turbine  46  and high-pressure turbine  54  rotationally drive respective low-speed spool  30  and high-speed spool  32  in response to the expansion. It will be appreciated that each of the positions of fan section  22 , compressor section  24 , combustor section  26 , turbine section  28 , and fan drive gear system  48  may be varied. For example, gear system  48  may be located aft of combustor section  26  or even aft of turbine section  28 , and fan section  22  may be positioned forward or aft of the location of gear section  48 . Each of compressor section  24  and turbine section  28  may include a rotor stage  100 . 
     With reference to  FIG. 2 , rotor stage  100  is shown. As will be appreciated by those skilled in the art, successive vanes  112  and rotor stages  100  are arranged serially along core flow path C. Vane  112  directs core airflow C as it traverses gas turbine engine  10  and toward downstream blade  102 . Downstream blade  102  extracts energy in the form of pressure from the core airflow C for application of rotational force to rotor disk  100  about engine axis A (shown in  FIG. 1 ). 
     Rotor stage  100  defines an interior cavity  108  and includes blade  102 , a rotor disk  104 , a disk cover  106 , and a spring damper  110 . Blade  102  has airfoil portion disposed within core flow path C and a root portion seated within rotor disk  104 . Disk cover  106  connects on its downstream side to rotor disk  104  by seating in a pocket defined on a forward face of rotor disk  104 . Disk cover  106  defines on its upstream side knife-edges that sealably couple with vane  112 . This separates hot gases traversing core gas path C from interior cavity  108  and allows for rotation of rotor disk  100  in relation to static engine components, e.g. vane  112 . 
     Spring damper  110  is disposed within interior cavity  108  and is attached to disk cover  106  such that such that blade  102 , rotor disk  104 , disk cover  106 , and spring damper  110  rotate with one of low-speed spool  30  (shown in  FIG. 1 ) and high-speed spool  32  (shown in  FIG. 1 ). As will be appreciated by those skilled in the art, disk cover  106  can be subject to forces during operation that can displace disk cover  106 , induce fatigue damage, or both, and therefore requires damping. Spring damper  110  is in intimate mechanical contact with disk cover  106  and provides a predetermined damping effect to disk cover  106  to counteract these forces. 
     With reference to  FIG. 3 , spring damper  110  is shown. Spring damper  110  has a split ring body  114 . Split ring body  114  defines a center  116  and a circular gap G separating a first end portion  118  and an opposed second end portion  120 . An evenly spaced segment  122  of split ring body  114  is evenly spaced with respect to center  116  and couples first end portion  118  to second end portion  120 . At least one of first end portion  118  and second end portion  120  is unevenly spaced from center  116  in relation to evenly spaced segment  122 . In this respect, at least one of first end portion  118  and second end portion  120  defines a curvilinear segment with a transition demarcated by a line tangent to an outer surface of evenly spaced segment  122 . As illustrated in  FIG. 3 , in certain embodiments, both first end portion  118  and second end portion  120  are unevenly spaced from center  116  with respect to evenly spaced segment  122 . 
     Evenly spaced body segment  122  spans a first angle A 1  in relation to center  116 . First end portion  118 , circular gap G, and second end portion  120  span a second angle A 2  in relation to the center  116 . Evenly spaced segment  122  is offset from center  116  by a substantially uniform offset (radial) distance along an arc spanning between about 180 degrees to about 270 degrees. In the embodiment illustrated, the arc spanned by evenly spaced segment  122  is about 270 degrees. Although evenly spaced segment  122  is illustrated in  FIG. 3  as a circular segment, it is to be understood that the shape and/or offset of spring damper  110  with respect to center  116  can be defined by a preload imposed by disk cover  106  (shown in  FIG. 2 ) as well as static and/or dynamic load(s) imposed on split ring body  114 . 
     Evenly spaced segment  122  is offset along by a first offset distance R 1  from center  116 . An end  124  of first end portion  118  is offset from center  116  by a second offset distance R 2  from center  116 . An end  126  of second end portion  120  is offset from center  116  by a third offset distance R 3  from center  116 . Second offset distance R 2  and third offset distances R 3  are greater than first offset distance R 1  such that first end  124  and second  126  are unevenly spaced outward from center  116  with respect to the evenly spaced segment  122 . 
       FIGS. 4-6  show exaggerated schematic views of spring damper  110  in an unloaded configuration, a statically loaded configuration, and a dynamically loaded configuration. In the unloaded configuration (shown in  FIG. 4 ), spring damper  110  is in a free state wherein substantially no force is applied to spring damper  110 . In the statically loaded configuration (shown in  FIG. 5 ), disk cover  106  imposes preload forces F on first end portion  118  and second end portion  120 . This orients first end portion  118  and second end portion  120  radially inwards, imparting a preload to the spring damper body and configuring spring damper  110  for resisting transportation loads. In the dynamically loaded configuration (shown in  FIG. 6 ), rotation of the assembly from operation exerts additional centrifugal force on spring damper  110 . This drives evenly spaced portion  122 , first end portion  118 , and second end portion  120  radially outward such that spring damper  110  has a substantially uniform radius. Arcuate segments defined by first end portion  118  and second end portion  122  as well as gap widths defined between first end  124  and second end  126  differ between each of the illustrated configurations. 
     With reference to  FIG. 4 , spring damper  110  is shown in the unloaded configuration. The unloaded configuration is a free state shape representative of an arrangement of spring damper  110  prior to installation into a circular device needing damping. Evenly spaced segment  122 , first end portion  118  and second end portion  120  collectively define an elliptical shape with a minor cord extending between 0 degrees (at the top of  FIG. 4 ) and 180 degrees (at bottom of  FIG. 4 ) and a major cord extending between 90 degrees (at left hand side of  FIG. 4 ), and the circumferential gap G. First end portion  118  and second end portion  120  define arcuate segments extending radially outwards from a circumference defined by an evenly spaced segment  122 . First end  124  and second end  126  are unevenly spaced in a radially outward arrangement in relation to center  116  and with respect to evenly spaced segment  122 . 
     With reference to  FIG. 5 , spring damper  110  is shown in a statically loaded configuration. The statically loaded configuration differs from the unloaded configuration in that spring damper  110  is installed in disk cover  106 . Disk cover  106  imposes preload forces F that cause spring damper  110  to have a smaller diameter relative to the unloaded configuration and which impart a preload that keeps the spring damper in place when subjected to transportation loads. As illustrated, spring damper  110  is mechanically connected to disk cover  106  (shown in dashed outline) to form an engine subassembly. As illustrated, disk cover  106  applies a preloading force F on first end portion  118  and second end portion  120  that orients first end portion  118  and second end portion  120  toward one another. This more directly aligns first end  124  with second end  126 . More direct alignment in turn causes tangentially oriented impacts, e.g. impact I, to cause first and second ends  124  and  126  to butt against one another, limiting reduction in the diameter of the part as a result of the event. This makes it more likely that spring damper  110  returns to its intended location following the event than split ring bodies with ends that overlay one another for a given transportation load. 
     With reference to  FIG. 6 , spring damper  110  is shown in a dynamically loaded configuration. The dynamically loaded configuration is similar to the statically loaded configuration with the addition of centrifugal forces associated with engine rotation R. Engine rotation R urges evenly spaced segment  122 , first end portion  124 , and second end portion  126  radially outward, further changing the arcuate shape of first end portion  124  and second end portion  126  and causing circumferential gap G to increase in width. As illustrated, width of circumferential gap G is wider in the dynamically loaded configuration than in the statically loaded configuration. Width of circumferential gap G is smaller than the width of circumferential gap G in the unloaded configuration. 
     As will also be appreciated by those skilled in the art, certain types of gas turbine engines and engine subassemblies can be subject to transportation loads while in a non-operating state. Transportation loads can exert forces on engine and/or engine subassembly sufficient to dislocate some types of damper from their intended location(s). Once dislocated, such dampers may be unable to provide an intended damping force (or effect) on engine structure requiring damping. 
     With respect to split ring dampers, Applicants have observed that transportation loads can sometimes be of sufficient magnitude to drive one end of a conventional split ring damper circumferentially past split ring body second end, causing one end of the damper to radially overlay another, and allowing the damper to dislocate from its intended position in relation to a structure requiring damping. Since dislocation can render the damper unable to provide its intended damping effect and/or potentially damage the engine operation, embodiments of the spring dampers described herein can provide greater resistance to dislocation due to tendency of the spring damper ends to remain in-plane with one another. This causes the opposed ends of the split ring body to butt against one another instead of overlap as result of the transportation load, making it more likely that the spring damper will return to its installed position rather than dislocate in response to the transportation load. This can be particularly advantageous when the transportation loads exert force tangent to the circumferential gap (as shown in  FIG. 5 ). 
     Embodiments of spring dampers described herein have end portions that are unevenly spaced when in their unloaded configuration. When installed in an engine or engine subassembly, these ends align with one another due to preloading force applied by the disk cover to the spring damper. This aligned causes a force associated with a transportation load to drive the end portion ends into contact with one another instead of overlap, limiting end portion displacement and making it more likely that the spring damper returns to its intended position rather than become dislocated. In embodiments, the end portion spacing increases the preload in the gap region when installed in a rotor disk assembly and makes the spring damper more resistant to dislocation. It can also cause the gap to be smaller when installed in a disk cover, potentially increasing the likelihood that the spring damper will remain in its intended position when subjected to transportation loads. It can further enable more favorable stress distribution in the spring damper, potentially reducing creep or other effects that could otherwise result in loss of preload caused by larger free state (uninstalled) diameter. 
     In embodiments, the split ring damper body transitions to a larger radius of curvature in the region of the ring gap (i.e. the circular gap). In certain embodiments, the transition is in the 0 degree to 180 degree of the ring body. This can increase the preload in the gap region when the split ring body in the vicinity of the circumferential gap when installed. It can also cause the gap to be smaller when installed in a disk cover or other circular structure. This smaller gap can further increase the ability of the ring to remain in its intended location when subjected to transportation loads. Further, it can enable a more favorable stress distribution in the split ring body, potentially allowing for customization of the split ring body to prevent creep related loss in preload otherwise caused by a larger free state diameter. 
     The apparatus, systems and methods of the present disclosure, as described above and shown in the drawings, provide for spring dampers with superior properties including improved resistance to dislocation due to transportation loadings or impact. While the apparatus and methods of the subject disclosure have been shown and described with reference to preferred embodiments, those skilled in the art will readily appreciate that changes and/or modifications may be made thereto without departing from the spirit and scope of the subject disclosure.