Abstract:
A gas turbine engine rotor ( 14 ), including: a rotor disk ( 16 ) comprising a bolt hole ( 82 ) there through and a counterbore ( 86 ); a bolt ( 18 ) configured to fit in the bolt hole and when so disposed to define an end ( 72 ) protruding beyond the counterbore; and a nut ( 84 ) configured to be disposed in the counterbore and to engage the protruding end. The counterbore is configured to permit limited eccentricity between the nut and the protruding end.

Description:
FIELD OF THE INVENTION 
     The present invention relates generally to the field of power generation, and more particularly to a gas turbine engine rotor, and specifically to a turbine rotor having a nut and bolt arrangement subjected to centrifugal forces acting orthogonal to a fastening direction. 
     BACKGROUND OF THE INVENTION 
     Gas turbine rotors may include several stages of rotor disks secured together with a stud (e.g. a bolt) and a nut. When spinning at high speed the weight of the nut results in a large amount of centrifugal force which must be reacted by the bolt threads. The centrifugal load can also impart an eccentricity between the bolt threads and the nut threads. Thus, the first engaging thread of the bolt, which bears most of the load in a standard nut and bolt configuration, experiences a circumferentially localized increase in load when spinning. If this effect is not, accounted for in the design of the nut and bolt it may reduce a life cycle of the nut and bolt arrangement. 
     Various attempts have been made to reduce the localized stress on the first engaging thread, many of which involve complicated manufacturing processes. Most of these are not specific to gas turbine engines. U.S. Pat. No. 8,038,377 to Ichiryu discloses a fastening device for a gas turbine engine rotor where the center of gravity of the nut is disposed in a nut hole (counterbore) and the nut is held concentric to the bolt to improve the axial and circumferential load distribution on the threads. However, in some configurations it is not possible to countersink the nut to this extent. Consequently, there remains room in the art for improvement. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The invention is explained in the following description in view of the drawings that show: 
         FIG. 1  is a sectional view of a gas turbine engine compressor showing a prior art nut and bolt arrangement. 
         FIG. 2  is a sectional view of the prior art nut and bolt arrangement of  FIG. 1  when experiencing centrifugal forces that occur during spinning of the rotor. 
         FIG. 3  is a sectional view of the nut and bolt arrangement disclosed herein. 
         FIG. 4  is a perspective view of the nut of  FIG. 4 . 
         FIG. 5  is a partial sectional view of the nut and bolt arrangement of  FIG. 4  when experiencing centrifugal forces that occur during spinning of the rotor. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     The present inventors have devised an innovative, inexpensive, and easy to manufacture nut and bolt arrangement that provides for a more even axial and circumferential distribution of loads on engaged threads when the nut and bolt are subjected to forces that act orthogonally to a fastening direction. The nut and bolt arrangement is particularly well suited for securing gas turbine engine rotor disks together when the nut cannot be substantially countersunk into the rotor. During operation the spinning of the rotor creates centrifugal forces that act on the nut and an end of the bolt that protrudes from the rotor disk. In a conventional and static nut and bolt arrangement the bolt&#39;s reactionary load is borne primarily by the first thread and the stress is essentially uniform circumferentially along the first engaging thread. In a spinning arrangement the centrifugal forces act to push the nut orthogonal to the fastening direction, causing an eccentricity between the bolt threads and the nut threads, and creating a bending moment on the protruding end of the bolt. The increase in force on the first engaging bolt thread, the change in direction of the force on the first engaging bolt thread, and a change in how the nut and bolt threads contact each other caused by the eccentricity all work together to locally increase stress on an inboard side (with respect to the axis of rotation of the rotor) of the first engaging bolt thread. The nut and bolt arrangement disclosed herein is configured to more evenly distribute the circumferential and axial stresses on the engaging threads. 
       FIG. 1  shows a portion of a compressor  10  of a gas turbine engine  12 , and a rotor  14  composed of several rotor disks  16  secured together with a prior art nut and bolt arrangement  18 . A prior art nut  20  is disposed on the bolt  22  at a cold end of the compressor  10 , while a downstream end  24  of the bolt is disposed closer to the turbine combustors  26 . The downstream end  24  may be secured by any means known to those in the art, including a second nut etc. During operation the rotor  14  rotates about a rotor axis of rotation  30  and this creates a centrifugal force  32  that urges the prior art nut  20  in a radially outward direction  34  (with respect to the rotor axis of rotation  30 ) orthogonal to a fastening direction  36 . This urging results in a minor radial (lateral) movement of the prior art nut  20  with respect to the bolt  22 , causing an eccentricity between the prior art nut  20  and the bolt  22 . When frictional forces between the prior art nut  20  and the bolt  22  are overcome, the radial movement is stopped when the prior art nut  20  and the bolt  22  bottom-out on each other. Bottoming out occurs when the prior art nut  20  can no longer move radially with respect to the bolt  22  due to their mechanical interaction without elastic or plastic deformation of the threads. For example, when flanks on both sides of one nut thread meet flanks on both sides of an associated bolt thread, or when a peak of the nut thread meets a valley of the bolt thread etc. 
       FIG. 2  is a representation of the prior art nut  20  and bolt  22  of  FIG. 1  when experiencing the centrifugal forces acting orthogonal to the fastening direction  36 . The prior art nut  20  has bottomed out against the bolt  22  due to a geometric relationship between a flank  40  on a first engaging thread  42  of the prior art nut  20 , a flank  40  on a first engaging thread  44  of the bolt  22 , and an interaction between a load face  46  of the prior art nut  20  and a load face  48  of the rotor disk  16 . As shown in  FIG. 2  and referred to herein, a top dead center (TDC)  60  and a bottom dead center (BDC)  62  of the nut are circumferential locations referring to a part of the prior art nut  20  farthest from the rotor axis of rotation  30  and a part of the prior art nut  20  nearest the rotor axis of rotation  30  respectively. At the BDC  62 , nut threads  64  are being forced into bolt threads  66 , causing a maximum lateral engagement locally between the threads. At the TDC  60 , nut threads  64  are separated from bolt threads  66 , causing a reduced lateral engagement locally between the threads. 
     The centrifugal forces present when the prior art nut  20  and bolt  22  are bottomed out cause a local concentration of high stresses at a high stress location  68  in the first engaging thread  44  of the bolt  22 . Since fatigue failures are the result of a material experiencing sufficient stress over time, and since the stresses in the high stress location  68  are relatively high, the fatigue life of the bolt  22  may be controlled by the fatigue life of the high stress location  68  in the prior art nut and bolt arrangement  18 . 
     As a result of the centrifugal forces and the wedging action resulting from a bottomed-out geometry, a crack  70  may initiate at the high stress location  68 , and may propagate throughout the bolt  22  as shown. The crack propagation may be aided by the nut  20  as the centrifugal forces essentially peel the prior art nut  20  around a tip  72  of the bolt  22  that protrudes beyond the load face  48  of the rotor disk  16 . A conventional bolt  22  may be, for example, up to ten feet long, or even longer, and may be up to 3.5 inches, or larger, in diameter. A liberated prior art nut  20  and portion  74  of the tip  72  therein thus represent considerable momentum and can cause damage within the gas turbine engine. Consequently, the inventors have devised the nut and bolt arrangement disclosed herein that is effective to better distribute the stresses circumferentially along each thread as well as among all the threads. This may increase the fatigue life of the nut and bolt arrangement. 
       FIG. 3  shows a nut and bolt arrangement  80  having a bolt  22  disposed in a bolt hole  82 , a nut  84  secured to the bolt  22  and partly disposed within a counterbore  86  in a rotor disk  16 . The nut  84  further includes a nut protruding end  90  that protrudes from the counterbore  86 , past an outer end  92  of the counterbore  86 . Both a nut face end  94  and the nut protruding end  90  may have cylindrical outer diameters to minimize the amount or material present in the nut  84  and the associated forces. As a result of the protrusion of the nut  84 , a center of gravity  96  of the nut  84  may rest outside the counterbore  86 . Radial nut holes  100  may be used to reduce weight and/or receive a tool (not shown) to enable assembly of the nut  84  onto the bolt  22 . As shown the radial nut holes  100  are through-holes. Alternately, they may be partial (blind) holes formed from the outer surface toward, but not reaching, the nut threads  134 . 
     Centrifugal forces  101  act on the center of gravity  96  of the nut  84  and create the nut bending moment  102  that the bolt  22  and associated bolt threads  66  must react. It is the radial centrifugal forces  101  and nut bending moment  102  and associated wedging action resulting from a bottomed-out geometry that cause the high stress location  68  in the prior art, but which is mitigated and/or eliminated using the nut and bolt arrangement  80  herein. 
     The nut face end  94  has a face end outer diameter  110 . The counterbore  86  has a counterbore inner diameter  112 . As shown the nut  84  is concentrically positioned within the counterbore  86  to form a fully concentric configuration  114  where a nut longitudinal axis  116 , a bolt longitudinal axis  118 , and a bolt hole longitudinal axis  120  are the same. A tolerance stacking between an outer diameter  130  of the bolt  22  and an inner diameter  132  of the bolt hole  82  may permit the bolt  22  to move laterally in the radially outward direction  34  with respect to the bolt hole  82 . A tolerance stacking between nut threads  134  and bolt threads  66  may permit the nut  84  to move laterally in the radially outward direction  34  with respect to the bolt  22 . If not laterally constrained, these tolerances would permit the nut  84  to move in the radially outward direction  34  until the nut  84  bottomed out into a bottomed out configuration  136  such as occurs in the prior art where there is no counterbore. 
     The counterbore inner diameter  112  used in the nut and bolt arrangement  80  is uniquely configured to be an optimized dimension that is larger than the face end outer diameter  110 . This is done to permit a limited amount of eccentricity, but to prevent a bottomed out configuration  136 . When the nut  84  moves radially/laterally and abuts a counterbore side wall  138 , the nut  84  reaches a maximum permitted eccentricity configuration  140 . The exact amount of a gap  126  desired between the face end outer diameter  110  and the counterbore inner diameter  112  when the nut and bolt arrangement  80  is in the fully concentric configuration  114  will depend on the tolerances between the nut  84  and the bolt  22 , the bolt  22  and the bolt hole  82 , the amount of deflection the tip  72  of the bolt  22  expected during operation resulting from the centrifugal forces on the tip  72  and the nut  84 , and the thread parameters etc. The gap may also account for a dilation (increase in diameter) of the nut  84  due to fastening forces that may axially com-press the nut. By controlling the amount of permitted eccentricity, the amount of contact area between the nut threads  134  and the bolt threads  66  at the TDC  60  and the nut threads  134  and the bolt threads  66  at the BDC  62  can be adjusted. Since stress is a result of force and area, adjusting the contact area permits the inventors to distribute the stress in the nut threads  134  and the bolt threads  66  circumferentially (from BDC  62  to TDC  60 ). 
     In addition to circumferential distribution control, the nut and bolt arrangement  80  permits axial distribution of the stresses via a unique undercut  150  which is annular in shape and surrounds the first engaging thread  42  of the nut  84  and up to three or even more nut threads  134 . The undercut  150  forms a unique conical section  152  within a nut counterbore  154  that includes a conical outer surface  156  that tapers inward at a taper angle  158  towards the nut threads  134 , and a face end  160  setback axially a distance from the load face  46  of the nut  84 . A nut counterbore inner surface  162  may meet the conical outer surface  156  and form a fillet  164  that extends around the circumference and hence has an annular shape. In the sectional view of  FIG. 3  the undercut  150  takes on a cantilevered shape. This cantilevered shape offers less structural support to the first engaging thread  42  of the nut  84  as well as less support to any such cantilevered nut threads  134  relative to other nut threads  134  not disposed on the conical section  152 . Consequently, any nut threads  134  on the conical section  152  are more readily axially and/or radially displaced. Permitting this axial displacement (via axial elastic deformation of the conical section  152 ) helps spread the forces and associated stresses from the first engaging thread  42  of the nut  84  to the adjacent nut threads  134 . This, in turn, helps spread the forces and stressed experienced by the associated bolt threads  66 . 
       FIG. 4  is a perspective view of the nut  84  showing the nut protruding end  90 , the nut face end  94 , the radial nut holes  100 , and the face end outer diameter  110 . 
       FIG. 5  is a partial sectional view of the nut and bolt arrangement  80  experiencing the centrifugal forces  101  that occur during spinning of the rotor  14  (not shown). As shown the nut and bolt arrangement  80  is in the maximum permitted eccentricity configuration  140 . The eccentricity can be seen by a difference in a BDC gap  166  between nut threads  134  and bolt threads  66  and a larger TDC gap  168  between nut threads  134  and bolt threads  66 . This means that some eccentricity between the bolt longitudinal axis  118  and the nut longitudinal axis  116 , and hence the bolt threads  66  and the nut threads  134 , has been permitted due to the selected counterbore inner diameter  112  designed to be larger than the face end outer diameter  110 . The tip  72  of the bolt  22  is essentially cantilevered from a remainder of the bolt  22  that is radially supported by the side wall of the bolt hole  82 . Consequently, while not visible in  FIG. 5 , the tip  72  of the bolt  22  may also be deflected radially outward (upward in  FIG. 5 ) due to the centrifugal forces acting on the tip  72  of the bolt. This may be exacerbated by the radial forces imparted on the tip  72  by the nut  84 . 
     The amount of eccentricity is selected based on the various factors mentioned above and this includes optimizing circumferential thread contact areas to account for competing factors. Specifically, a TDC contact area  170  and a BDC contact area  172  can be optimized to provide an amount of contact area that is responsive to the loads at the respective areas. While the TDC  60  and the BDC  62  are discussed herein for sake of clarity, the concepts apply to the entire circumference of the nut threads  134  and the bold threads  66 . 
     While rotating, at the BDC  62  the first engaging thread  42  of the bolt  22  must react an axial load  178  resulting from the fastening of the nut  84  with the bolt  22 , a radial/lateral load resulting from the centrifugal forces  101  on the nut  84 , and a bending moment load from the nut bending moment  102 . A resulting BDC load  174  on the first engaging thread  44  of the bolt  22  may be at a BDC angle  176  from parallel with the bolt longitudinal axis  118 . At the TDC  60  the first engaging thread  42  of the bolt  22  reacts with the axial load resulting from the fastening of the nut  84  with the bolt  22 , and perhaps with a negligible bending load from the nut bending moment  102 . 
     However, the centrifugal forces not only urge the nut  20  radially outward, but they also urge the cantilevered tip  72  of the bolt  22  radially outward (upward in  FIG. 5 ). The remainder of the bolt  22  within the bolt hole  82  maintains its radial position because it is held in place by the bolt hole  82 . As a result, the tip  72  of the bolt  22  can be envisioned as rotating slightly clockwise during operation as a result of its radially outward movement. This causes the bolt threads  66  at TDC  60  to shift slightly axially to the right, moving the bolt threads  66  toward the nut threads  134  at TDC  60 . The amount of axial shift increases with the distance a given thread is located from an effective pivot point (not shown). Consequently, there may be more axial shift in third and fourth threads than in the first engaging threads  42 ,  44 , for example. Simultaneously, as the nut  20  shifts, the nut threads  134  at TDC  60  disengage from the bolt threads  66 . In contrast, the bolt threads  66  at BDC  62  shift slightly axially to the left, away from the nut threads  134  at BDC  62 . Likewise, the amount of axial shift increases with the distance a given thread is located from an effective pivot point. Simultaneously, as the nut  20  shifts laterally, the nut threads  134  at BDC  62  increase engagement with the bolt threads  66 . 
     In an arrangement where the nut  20  is prevented from any lateral movement and the tip  72  of the bolt  22  protrudes, the axial shift of the bolt threads  66  at TDC  60  of a deflecting tip  72  of the bolt may cause stresses at the TDC  60  to be greater than at BDC  62 . In an arrangement where the nut  20  is unrestrained laterally and the tip  72  of the bolt  22  protrudes, the wedging effect of the bottomed out configuration  136  may cause the high stress location  68  at the BDC  62 . 
     The inventors have recognized that stress locations vary depending on the configuration, and the nut and bolt arrangement  80  disclosed herein falls between not enough eccentricity (high stress at TDC  60 ), and too much eccentricity (high stress at BDC  62 ). In particular, the inventors have recognized that stresses can be distributed by striking a balance between several factors associated with increasing eccentricity, including; increasing axial shift of the bolt threads  66  at TDC  60  toward the right due to tip  72  rotation (increasing force at TDC  60 ); disengagement of the nut threads  64  from the bolt threads  66  at TDC  60  due to lateral nut  20  movement (decreasing force at TDC); decreasing TDC contact area  170  (tending to increase stress at TDC  60 ); increasing axial shift of the bolt threads  66  at BDC  62  to the left due to tip  72  rotation (decreasing force at BDC  62 ); deeper engagement of the nut threads  64  with the bolt threads  66  at BDC  62  due to lateral nut  20  movement (increasing force at BDC  62 ); increasing BDC contact area  172  (tending to decrease stress at BDC  62 ); and preventing the bottomed out configuration  136 . Consequently, the inventors have recognized that by permitting a limited eccentricity they can tailor the TDC contact area  170  and the BDC contact area  172  of a particular nut and bolt arrangement  80  having certain parameters to match the magnitude of their respective loads on the bolt threads  66  for an expected set of operating conditions. The circumferential stress distribution permitted by permitting the limited eccentricity, together with the axial stress distribution permitted by the undercut  150 , provide for much more evenly circumferentially and axially distributed stresses on the bolt threads  66 . This, in turn, may extend the fatigue life of the bolt threads  66 . 
     While in the maximum permitted eccentricity configuration  140  the nut threads  134  are also experiencing loads and stresses, as is the conical section  152 . In one model, a peak axial load on the first engaging thread  42  of the bolt  22  was reduced by approximately fifty percent from the prior art. This load was transferred to the other bolt threads, such that the first engaging thread  42  of the bolt  22  experienced a load that was less than a load experienced by many adjacent bolt threads  66 . In that model the load on the third and fourth threads was among the greatest of the bolt threads  66 . This redistribution is transferred to the nut threads  134 . In order to keep stress in the conical section  152  more uniform, an area of the conical section  152  associated with each nut thread  134  is also tailored to match the load of the associated nut thread  134 . For example, a cross section  1 - 1  of the conical section  152 , taken orthogonal to the nut longitudinal axis  116  and at a nut thread major diameter  180 , is characterized by a  1 - 1  cross sectional area. Likewise, cone cross section  2 - 2 ,  3 - 3 , and  4 - 4  are associated with respective threads and are characterized by respective cross sectional areas. 
     Since the load on the first engaging thread  42  of the nut  84  is relatively low, the cross sectional area of cross section  1 - 1  may be relatively low. On the other hand, since the load on a second nut thread  182 , a third nut thread  184 , and a fourth nut thread  186  increases, the respective cross sectional areas of sections  2 - 2 ,  3 - 3 , and  4 - 4  may be larger to accommodate their respective greater loads. The cross sectional area of the various sections can be controlled by controlling the taper angle  158  of the conical outer surface  156  and a location of the conical outer surface. In this manner the stresses in the conical section  152  may be more evenly axially distributed. While it may be possible to circumferentially vary the shapes of the sections  1 - 1 ,  2 - 2 ,  3 - 3 , and  4 - 4  to accommodate the circumferential variation in load, such as by permitting an eccentricity between an inner diameter and an outer diameter of the cross section, this may require burdensome machining and assembling methods. 
     From the foregoing it can be seen that the inventors have devised a clever, unique, and yet simple solution that can turn more uniformly distribute stresses in among the threads in a rotating nut and bolt arrangement that experiences uneven loads on the threads. Consequently, this represents an improvement in the art. 
     While various embodiments of the present invention have been shown and described herein, it will be obvious that such embodiments are provided by way of example only. Numerous variations, changes and substitutions may be made without departing from the invention herein. Accordingly, it is intended that the invention be limited only by the spirit and scope of the appended claims.