Abstract:
A seal assembly is provided for use in a hydraulic actuator of the type including a cylinder, a piston movably disposed in the cylinder, and a piston rod connected to the piston and extending from an end of the cylinder. The seal assembly comprises a seal member formed of a material impervious to hydraulic fluid and having an inner edge and an outer peripheral edge. The inner edge of the seal member defines a central opening which is larger than the outside diameter of the piston rod for receiving the rod. The outer edges of the seal member are attached adjacent the end of the cylinder so that the inner edge of the seal member is radially spaced from the outside diameter of the piston rod and the seal member forms a fluid impervious wall for preventing hydraulic fluid from contacting the elastomeric portion of the bearing assembly.

Description:
CROSS-REFERENCES 
     This application claims the benefit of U.S. Provisional Application No. 60/233,308, filed Sep. 15, 2000, the contents of which are hereby incorporated by reference. 
    
    
     GOVERNMENT RIGHTS 
     The Government has rights to the invention pursuant to government contract N 000014-96 -C-2079 awarded by the United States Naval Research Laboratory. 
    
    
     BACKGROUND 
     The present invention relates generally to a hydraulic actuator and, more particularly, to an improved hydraulic actuator for use in an active mount for a vibrating component in a system for reducing vibration and noise transmission from the vibrating component to a support structure. 
     Hydraulic actuators are used in numerous environments to induce movement of one object with respect to another object. A hydraulic actuator generally includes a cylinder and a moveable piston inside the cylinder. A piston rod is connected to the piston and extends outwardly from one end of the cylinder where the rod end is attached to the first object. The other end of the cylinder is mounted, directly or indirectly, to the second object. The means for mounting the piston rod and cylinder to the objects may incorporate flexible bearing assemblies to provide some “softness” to the attachment. Such bearing assemblies preferably comprise elastomeric material. Pressurized hydraulic fluid is introduced into the interior of the cylinder on one or both sides of the piston to effect longitudinal movement of the piston in the cylinder so that the objects are moved relative to one another. 
     Hydraulic actuators may be used as a component of an active mount in a system for reducing vibration and noise transmission from a vibrating component to a support structure. For example, hydraulic actuation systems are used for actively reducing the vibratory and acoustic loads on aircraft, particularly rotary wing aircraft such as helicopters. A primary source of vibratory and acoustic loads in a helicopter is the main rotor system. The main rotor system of a helicopter includes rotor blades mounted on a vertical shaft that projects from a transmission, often referred to as a gearbox. The gearbox comprises a number of gears which reduce the rotational speed of the helicopter&#39;s engine to the much slower rotational speed of the main rotor blades. The gearbox has a plurality of mounting “feet” which are connected directly to structure in the airframe which supports the gearbox. The main rotor lift and driving torque produce reaction forces and moments on the gearbox. All of the lift and maneuvering loads are passed from the main rotor blades to the airframe through the mechanical connection between the gearbox feet and the airframe. The airframe structure which supports the gearbox is designed to react to these primary flight loads and safely and efficiently transmit the flight loads to the airframe. 
     In addition to the nearly static primary flight loads, the aircraft is also subjected to vibratory loads originating from the main rotor blades and acoustic loads generated by clashing of the main rotor transmission gears. The vibratory and acoustic loads produce vibrations and audible noise that are communicated directly to the helicopter airframe via the mechanical connection between the gearbox and the airframe. This mechanical connection thus becomes the “entry point” for the unwanted vibration and noise energy into the helicopter cabin. The vibrations and noise within the aircraft cabin cause discomfort to the passengers and crew. In addition, low frequency rotor vibrations are a primary cause of maintenance problems in helicopters. 
     Active vibration and noise reduction systems in aircraft utilize sensors to monitor the status of the aircraft, or the vibration producing component, and a computer-based controller to command actuators to reduce the vibration and noise. The sensors are located throughout the aircraft and provide signals to the adaptive controller. The controller provides signals to the hydraulic actuation system, including a plurality of actuators located at strategic places within the aircraft. The actuators produce controlled forces or displacements which attempt to minimize vibration and noise at the sensed locations. 
     Two methods of actuator placement are frequently used in the active system: (1) distribution of actuators over the airframe, or (2) co-location of the actuators at, or near, the vibration or noise entry point. When applied to the main rotor system of a helicopter, the co-location approach places the actuators at or near the structural interface between the transmission and airframe stopping vibration and noise near the entry point before it is able to spread out into the aircraft. This has the advantage of reducing the number of actuators and the complexity of the control system. Active systems using this approach to counteract vibration and noise employ actuators mounted in parallel with the entry point or in series between the transmission gearbox feet and airframe support structure. 
     When the actuator is mounted in series with the vibrating component and the support structure, the point of attachment of the piston rod has six possible degrees of freedom. However, only the degree of freedom along the principle load carrying axis can be actively controlled for vibration and noise reduction. Further, since the elastomeric bearing is located between the piston rod and the attachment point, the bearing assembly must provide high static and dynamic stiffness along this active, load carrying axis so that motions of the piston translate directly into unattenuated motions at the attachment point. To ensure that motions at the attachment point along the five non-active degrees of freedom do not create vibration and noise, the stiffness between the attachment point and actuator along these directions must be low. However, the need for the elastomeric bearing to be stiff along its principle load carrying axis yet soft about the other five degrees of freedom can cause the elastomeric bearing to be unstable under load. Also, transverse motions and rotational forces on the piston can induce high loads between the piston and cylinder which may cause the piston to bind. 
     Moreover, since hydraulic actuators operate under high pressure, the leakage of hydraulic fluid often occurs, particularly where the piston rod passes through the end of the cylinder. This leads to maintenance problems as well as environmental concerns. Additionally, escaping hydraulic fluid can damage the elastomeric material of the bearing assembly. Cylinder seals are usually provided to prevent leakage of hydraulic fluid from the cylinder. Cylinder seals are typically flat, disc-shaped bodies having a central opening. The seals peripherally engage the cylinder body at an outer edge and receive the piston rod through the central opening. The cylinder seal allows movement of the rod relative to the cylinder while at the same time sealing the interface between the rod and cylinder to block fluid leakage along the rod. However, cylinder seals have been known to wear out frequently, primarily due to friction between the rod and seal which results in deterioration of the sealing function and eventual leakage. When the seal fails, replacement requires the removal of the entire hydraulic actuator from service. 
     For the foregoing reasons, there is a need for an active mount including a hydraulic actuator comprising an elastomeric bearing which is stable under load. The new hydraulic actuator should withstand the significant loads generated when used in an active mount for the transmission of the main rotor system of a helicopter. Further, any transverse or rotational forces on the actuator should not induce high loads between the piston and the housing. The new hydraulic actuator should also have improved means for preventing leakage of hydraulic fluid from the cylinder. The leakage preventing means should protect the elastomeric components of the bearing assembly from damage by hydraulic fluid. The leakage prevention means should also have excellent wear characteristics. 
     SUMMARY 
     Therefore, it is an object of the present invention to provide a hydraulic actuator including an elastomeric bearing that remains stable under load and does not induce high loads between the piston and cylinder due to transverse motions and rotations at the point of attachment. 
     It is also an object of the present invention to provide a hydraulic actuator which prevents leaked hydraulic fluid from contacting the elastomeric bearing assembly. 
     According to the present invention, a seal assembly is provided for use in a hydraulic actuator of the type including a cylinder, a piston movably disposed in the cylinder, and a piston rod connected to the piston and extending from an end of the cylinder. The seal assembly comprises a seal member formed of a material impervious to hydraulic fluid and having an inner edge and an outer peripheral edge. The inner edge of the seal member defines a central opening which is larger than the outside diameter of the piston rod for receiving the rod. Means are provided for sealing the inner and outer edges of the seal member adjacent the end of the cylinder so that the inner edge of the seal member is radially spaced from the outside diameter of the piston rod and the seal member forms a fluid impervious wall. A bearing assembly may be connected to the end of the piston rod externally of the actuator cylinder and the sealing means may comprise first and second rings removably attached to the bearing assembly so that the inner and outer edges, respectively, of the seal member are sealingly secured between the rings and the bearing assembly for preventing hydraulic fluid from contacting the elastomeric portion of the bearing assembly. Preferably, the seal member material is flexible for accommodating relative movement between the piston and the cylinder. 
     Also according to the present invention, an active mount is provided for mounting a vibrating component to a support structure, such as a gearbox to the airframe of an aircraft, for reducing vibration transmission from the vibrating component to the support structure. The active mount comprises a housing adapted to be attached to one of the vibrating component or the support structure and a hydraulic actuator disposed in the housing. The hydraulic actuator includes a cylinder, a piston movably disposed in the cylinder, a piston rod connected to the piston and extending outwardly of an end of the cylinder and connected to the other of the vibrating component or the support structure. A seal member formed of a material impervious to hydraulic fluid and having a central opening which is larger than the outside diameter of the rod for receiving the rod is sealed in the cylinder so that the inner edge of the seal member is radially spaced from the outside diameter of the piston rod and the seal member forms a fluid impervious wall. A bearing assembly may be connected between the second end of the piston rod and the other of the vibrating component or the support structure and the sealing means may comprise first and second rings removably attached to the bearing assembly so that the inner and outer edges, respectively, of the seal member are sealingly secured between the rings and the bearing assembly for preventing hydraulic fluid from contacting the bearing assembly. Preferably, the seal member material is flexible for accommodating relative movement between the piston and the cylinder. 
     According to another aspect of the present invention, a hydraulic actuator is provided for an active vibration and noise control system in an aircraft, the actuator including a mounting member for attaching to the vibrating component and a bearing assembly engaged with mounting member and the piston. The bearing assembly comprises an inner bearing member, a first bearing located between the mounting member and the inner bearing member for transmitting axial motions between the mounting member and the piston, a second bearing located between the inner bearing member and the housing for transmitting moment and shear loads from the mounting member to the housing, and a third bearing located between the inner bearing member and the piston for permitting rotational movement of the inner bearing member relative to the piston. The bearings of the bearing assembly are located so that their central axes are substantially aligned. 
     The coincidence of the central axes of rotation of the several bearings provides a bearing assembly which is stable under load. This arrangement also significantly reduces the force which is transmitted between the piston and housing when the elastomeric bearing is loaded, which must be kept low to ensure the piston does not bind in its bore. The coincidence of the centers of rotation allows the use of a spherical bearing as the third bearing with a very low rotational stiffness. Hence rotation of the radial journal bearing relative to the spherical bearing creates low torques between the two which in turn creates low reaction forces between the piston and housing. The sealing assembly prevents leaked hydraulic fluid from reaching, and damaging, the elastomeric bearing. 
     The foregoing and other features and advantages of the present invention will become more apparent in light of the following detailed description of the preferred embodiments thereof, as illustrated in the accompanying figures. As will be realized, the invention is capable of modifications in various respects, all without departing from the invention. Accordingly, the drawings and the description are to be regarded as illustrative in nature, and not as restrictive. 
    
    
     DRAWINGS 
     For the purpose of illustrating the invention, the drawings show a form of the invention which is presently preferred. However, it should be understood that this invention is not limited to the precise arrangements and instrumentalities shown in the drawings. 
     FIG. 1 is a schematic representation of a helicopter transmission arrangement incorporating the present invention. 
     FIG. 2 is a schematic representation of an embodiment of a system according to the present invention for reducing vibration and noise passing from a helicopter transmission gearbox to the airframe. 
     FIG. 3 is graph showing eigenvalues for static forces on a helicopter transmission gearbox foot. 
     FIGS. 4-6 are graphs showing the eigenforces for eigen value 12 shown in FIG.  3 . 
     FIGS. 7-9 are schematic representations of the orientation directions for actuators in an active helicopter transmission mount according to the present invention. 
     FIG. 10 is a schematic representation of a hydraulic actuation system according to the present invention for delivering fluid to a plurality of actuators. 
     FIG. 11 is a schematic representation of an active transmission mount connected to an aircraft frame and transmission gearbox foot according to the present invention. 
     FIG. 12 is an exploded view of an active transmission mount connected to an aircraft frame and transmission gearbox foot according to the present invention. 
     FIG. 12 a  is an exploded view of the transmission gearbox foot and foot attachment for an active transmission mount shown in FIG.  12 . 
     FIG. 13 is a cutaway perspective view of the hydraulic actuation system for a gearbox mounting foot. 
     FIG. 14 is side elevation view of the hydraulic actuation system shown in FIG.  5 . 
     FIG. 15 is an isometric view of one embodiment of an actuator according to the present invention. 
     FIG. 16 is a cross-sectional view of the actuator shown in FIG. 16 taken along line  16 — 16 . 
     FIG. 17 is an exploded view of the actuator shown in FIG.  15 . 
     FIG. 18 is an enlarged view of a portion of the actuator shown in FIG.  16 . 
     FIG. 19 is an exploded view of the bearing assembly shown in FIG.  17 . 
     FIG. 20 is a close-up cross-sectional view of a tuned stub assembly. 
     FIG. 21 is a close-up cross-sectional view of a tuned stub with an adjustment screw assembly. 
    
    
     DESCRIPTION 
     While the invention will be described in connection with one or more preferred embodiments, it will be understood that it is not intended to limit the invention to those embodiments. On the contrary, it is intended that the invention cover all alternatives, modifications and equivalents as may be included within the spirit and scope of the invention as defined by the appended claims. 
     Certain terminology is used herein for convenience only and is not to be taken as a limitation on the invention. For example, words such as “upper,” “lower,” “left,” “right,” “horizontal,” “vertical,” “upward,” and “downward” merely describe the configuration shown in the Figures. Indeed, the components may be oriented in any direction and the terminology, therefore, should be understood as encompassing such variations unless specified otherwise. 
     Referring now to the drawings, wherein like reference numerals illustrate corresponding or similar elements throughout the several views, FIG. 1 illustrates a transmission arrangement  20  for a helicopter. The transmission arrangement  20  includes a gearbox  22  which is connected to a helicopter rotor head (not shown). The gearbox  22  is also connected to the drive train  24  of the helicopter&#39;s engine  26 . The gearbox  22  is supported by an airframe comprising a structural element  28 . The gearbox  22  includes a plurality of mounting feet  30  which are attached to the airframe structure  28 . In accordance with the present invention, active transmission mounts (ATMs)  32  are mounted in series between the each gearbox mounting foot  30  and the airframe structure  28  for isolating the mounting feet  30  of the main rotor gearbox  22  from the airframe. 
     According to the present invention, the ATM  32  is a part of an active transmission mount system  34 , an embodiment of which is schematically illustrated in FIG.  2 . The ATM system  34  comprises one or more hydraulic ATM actuators  36  associated with each of four ATMs  32 , a plurality of sensors  38 ,  40 ,  42  positioned throughout the aircraft, an electronic controller  44  which sends signals to a hydraulic actuation system  46  for commanding the actuation system to actuate the ATM actuators  36  according to the desired operational state. 
     The sensors comprise position sensors  38  for monitoring the static position of the feet  30  relative to the airframe  28 . These sensors  38  are used along with the controller  44  and the actuation system  46  to ensure that the transmission does not move out of static alignment with other elements of the airframe. The preferred location and type of the other sensors  40 ,  42  are a function of the type of control approach used by the controller  44 . For example, one type of control approach utilizes sensors  40  that are located adjacent to the mounting feet  30  and the ATM&#39;s  32 . These sensors  40  comprise accelerometers, to sense airframe acceleration. This same control approach may use pressure sensors  42  to sense dynamic pressure fluctuations in the actuator fluid lines  50 . An alternate control approach may use accelerometers  40  mounted at selected locations within the airframe, such as at the foot of the pilot or a seat. The choice of local sensors (accelerometers  40  or pressure sensors  42 ) or remote accelerometers  40  is largely based on the type of airframe to which the ATM system  34  is applied and is also based on the stiffness requirements defined for the ATM actuators  36 . 
     The signals output from the sensors  38 ,  40 ,  42  are provided for processing to the controller  44  which comprises a signal processor, computer, or the like. For each mounting foot  30 , the controller  44  determines the position of the foot  30  and vibratory loading of the airframe based on the sensed signals being transmitted by the sensors. The controller  44  then determines a desired operational state for each ATM actuator  36  as a function of one or more of the sensed signals and operates to nullify position offset of the gearbox  22  while also reducing the vibratory load passing through the ATMs  32  and into the airframe. 
     The hydraulic actuation system  46  supplies a hydraulic fluid under pressure to each hydraulic actuator  36  so that the actuator moves in the desired manner and at the desired frequency to reduce the sensed vibrations emanating from a mounting foot  30  of the gearbox  22  passing into the airframe. In the illustrated embodiment, the hydraulic actuation system  46  includes one or more electro-hydraulic valves  48  which are each electrically connected to the controller  46  via a control line  52  for supplying current to the valve  54 . For example, the hydraulic inputs of the two actuators  36  shown in FIG. 2 are interconnected into a common hydraulic fluid line  50  and connected to the hydraulic control valve  48 . For the sake of simplicity, only a single hydraulic control valve  48  and associated hydraulic interconnections are shown. The controller  44  generates output control signals to the hydraulic control valve(s)  48  in response to the signals received from the sensors  38 ,  40 ,  42 . The valve(s)  48  open and close in response to the output control signals to provide a vibratory flow of high pressure hydraulic fluid from a fluid source (not shown) to and from the actuators  36 . In FIG. 2, the supply flow into the valve  48  is generally indicated at  54  and the flow to the ATM actuators  36  is generally indicated at  56 . The hydraulic pressure and location of the actuators&#39; pistons are thus adjusted by the controller  44  based on the signals from the sensors  38 ,  40 ,  42 . 
     The active transmission mount system  34  of the present invention acts to isolate the vibratory and acoustic loads generated by the main rotor gearbox  22  from the airframe. The ATM system  34  achieves vibration reduction by controlling the applied fluid flow within the ATM actuators  36 , and thus the hydraulic pressure acting on the pistons in the actuators  36 . A quasi-steady pressure is applied to each actuator  36  to react to the applied quasi-steady flight and maneuvering loads. The vibratory loads that are applied along the actuator&#39;s principle, or “active”, axis are transmitted into the hydraulic column. This causes cancellation of pressure fluctuations which would otherwise be transmitted into the airframe causing vibration if left unaltered. Generally, an increase in hydraulic pressure on the pistons when a vibratory load pushes on the actuator  36  is relieved by the ATM system  34  by removing fluid, and a decrease in hydraulic pressure when a vibratory load pulls on the actuator  36  is accommodated by the ATM system  34  by increasing hydraulic fluid flow to the actuator. Hence, the actuator  36  is operated by removing and supplying a sufficient amount of hydraulic fluid against the head of the piston to allow the piston to translate in substantially the same direction and at substantially the same frequency as the vibrating gearbox  22 . In this way, the ATM system  34  allows relative motion between the gearbox  22  and the airframe at low vibration frequencies, typically greater than about 2 Hz, so that the gearbox  22 , in effect, floats in a dynamic sense with respect to the airframe, but maintains a steady, static position relative to the airframe. As a result, vibratory pressure is minimized, thereby reducing the transfer of vibration related to the applied rotor vibratory loads from the ATM  32  to the airframe. 
     Also, as seen in FIG. 2, the actuation system  46  preferably includes a passive noise isolator  58 . The passive isolator  58  introduces softness into the hydraulic system at predetermined frequencies to allow the system to attenuate high frequency (e.g., &gt;500 Hz) and low amplitude, 1/1000 inch, noise that is otherwise transmitted by the gearbox feet  30  to the ATM  32  causing high frequency noise in the fluid lines  50  which, in turn, leads to noise in the aircraft. In order to reduce this high frequency noise, the hydraulic line  50  is connected to the passive isolator  58 . It is understood that the other hydraulic lines (not shown) that interconnect the valves  48  and their associated actuators  36  are also connected to passive isolators. 
     According to the present invention, a system analysis is performed to determine the ATM actuator  36  placement, orientation and size (piston diameter). The actuator  36  placement, orientation and size is determined principally by the static loads between the transmission  22  and the airframe needed to keep the relative position of the transmission and the airframe constant at maneuvering frequencies, which are effectively static. The actuators  36  must supply these loads under all operating conditions. In addition, the actuators must respond to aircraft dynamic loads. Preferably, the actuators  36  should be no larger than necessary, as large actuators result in added weight and power consumption. The actuator configuration selected should use a minimum number of actuators to minimize system cost and maximize system reliability. And, to simplify actuator design, the actuators selected should use only single acting hydraulic pistons. 
     The system analysis of the present invention is an iterative process involving the following steps: (1) development of a mathematical model of the relative movement of the transmission and the airframe at the transmission mounting locations based on finite element models (FEM) of the transmission and airframe, (2) application of a sequence of flight scenarios, or cases, that define the motion of the transmission at a variety of load and speed conditions, (3) analysis of the motion of the transmission feet for each of the flight cases, (4) determination of an actuator placement and orientation scheme which satisfies all of the flight scenarios, and (5) iteration of step (4) until actuator sizing and orientation lead to satisfactory performance of the actuators under the loads described by steps (1)-(3), and taking into account any actuator physical mounting constraints for the particular application. 
     An example of the application of this analysis was performed on Sikorsky Aircraft Corporation&#39;s S-76® aircraft (S-76® is a registered trademark of the Sikorsky Aircraft Corporation). Sikorsky Aircraft Corporation has developed a 60,000 degree of freedom FEM of the transmission and an 11,000 degree of freedom FEM of the airframe for the S-76®. Using the FEM and well-known mathematical operations, a model was constructed which defines the motion of specific grid points in the FEM. This is a standard use of FEM modal outputs to construct a simplified mathematical model of the structure of interest. For example, for the ATM vibration response, only low frequency modes, those with modal frequencies below 200 Hz, need be used to construct a response function describing the structure of interest. This mathematical response function can then be used to determine the resulting frequency response of the structure to a driving force over the frequency range of interest. 
     For the ATM system, the grid points of interest are the locations where the transmission gearbox  22  mounts to the airframe. Forces at the grid points representing the main rotor hub intersection with the transmission are provided by a simulation of helicopter response. Sikorsky Aircraft Corporation developed a simulation of the S-76® aircraft&#39;s performance called GENHEL. The GENHEL simulation allows the user to determine static loads at the rotor hub based on specific aircraft loads and aircraft speeds. This simulation was used to produce static hub loads at aircraft loads of 8,750 lbs, 11,700 lbs and 12,800 lbs. For each load case, aircraft speeds were selected at 20 mph increments beginning at near zero (hover) up to 160 mph, for a total of 9 cases per load. In addition, the 8,750 lb and 11,700 lb load cases were also run at 180 mph for a total of 29 combinations of weight and speed. These 29 cases were the flight scenarios used to design the ATM system for the S-76® aircraft. 
     To analyze the motions occurring at the transmission mounting locations the mathematical models of the transmission gearbox mounting foot  30  motion and the airframe motion are linked together at the foot locations using the mathematical equivalent of “soft” springs. Using the “soft” springs allows all of the motion between the foot and the airframe to be accounted for by motion in the springs. The springs are used to calculate the supporting forces the actuator  36  would have to supply. The following sequence of equations shows the structure of the combined transmission and airframe mathematical model.                [     x   1     ]     n     =         ∑     m   =   1     12                    [     A   1     ]       n   ,   m            [     F   1     ]       m          
     [     x   2     ]     n       =         ∑     m   =   1     12                [     A   2     ]       n   ,   m            [     F   1     ]       m       +       ∑     m   =   1     6                [     A   3     ]       n   ,   m            [     F   3     ]       m                  
                        [     F   1     ]     n     =       ∑     m   =   1     12                [   B   ]       n   ,   m            [       x   1     -     x   2       ]       m                                  
     where m=1, . . . , 12, 
     A 1  is the 12×12 airframe compliance at the 4 mounting transmission feet mounting locations, excluding torques, 
     A 2  is the 12×12 airframe compliance matrix at the bottoms of the 4 transmission feet, 
     A 3  is the 12×6 transfer compliance matrix between the 3 forces and 3 torques at the rotor hub and the 12 linear forces at the 4 transmission feet, 
     B is a 12×12 diagonal stiffness matrix for 12 fictitious springs between the 4 transmission feet and the 4 airframe pads, 
     [X 1 ] n , where n=1, . . . , 12, is the airframe displacement at each of the four mounting locations in each of 3 orthogonal directions, 
     [X 2 ] n , where n=1, . . . , 12, is the airframe displacement at each of the four mounting locations in each of 3 orthogonal directions, 
     [F 3 ] m , where m=1, . . . , 6, is a 6 degree-of-freedom force vector applied to the transmission at the intersection of the main rotor hub and transmission, 
     [x 1 −x 2 ] m , where m=1, . . . , 12, is the spring displacement at each of four mounting locations in each of 3 orthogonal directions, and 
     [F 1 ] m , where m=1, . . . , 12, is the resultant force applied to the base of the transmission at the airframe mounting pads from the force vector [F 3 ]. 
     The motion at the transmission mounting location is represented by the differential motion between the airframe motion at the mounting location, x 1 , and the motion at the base of the transmission, x 2 . The differential motion, x 1 −x 2 , when applied to the spring stiffness of the “soft” springs produces a disturbance force, F 1 , which is applied at both the airframe and the transmission. The model is driven by hub forces from each of the 29 flight scenarios described above indicated by F 3 . The elements in the compliance matrices, A 1 , A 2  and A 3  are constructed from the simplified mathematical models of the transmission and airframe described above. 
     The next step in the system analysis is determination of the force vectors, F 1 , resulting from each of the 29 flight scenarios described above. For each flight scenario, the GENHEL simulation provides a force vector, F 3 , of hub forces. The forces between the airframe and the transmission are assumed to be completely described by a three-dimensional force vector. Since the S-76® transmission has four feet, there are a total of twelve components of force for the S-76® aircraft model. Further, since the transmission is a rigid body at these frequencies and is limited to six degrees of freedom, only six linearly independent combinations can be formed of these twelve components. A symmetric, 12×12 covariance matrix is constructed by forming the outer product of the forces F 1  resulting from each of the 29 flight cases. Each element in the covariance matrix is then summed over the set of 29 cases:          C     n                 m       =       ∑     α   -   1     29            F   n   α          F   m   α                                
     where 
     alpha is the aircraft flight test case from 1 to 29, and 
     F n  and F m  are forces in the x, y, and z directions at each of the four mounting feet, so n and m range from 1 to 12. 
     Because the transmission is limited to six degrees of freedom, the covariance matrix has six nonzero eigenvalues and six zero eigenvalues. An eigenvalue is the mean square force required for a support mode. FIG. 3 shows the values of each of the 12 eigenvalues for the S-76® aircraft application ordered from smallest to largest. The eigenvalues and eigenvectors resulting from the diagonalization of the force covariance matrix described above are determined using standard matrix solution software. The largest eigenvalue, number 12, has a magnitude approximately 30 times that of eigenvalues 10 and 11. Eigenvalues 10 and 11 are approximately 100 times larger than eigenvalue 9 and so on. Eigenvalues 1-6 are essentially zero. The ATM solution for the S-76® was thus focused on the three largest eigenvalues, eigenvalues 10-12, since the remaining eigenvalues have a significantly smaller impact on system response. Eigenvectors corresponding to each of the eigenvalues are the forces needed for the transmission to support the airframe. Examination of the forces corresponding to each eigenvector reveals its physical origin. For example, the eigenvector corresponding to eigenvalue number 12 (FIGS. 4-6) is a combination of vertical lift and rotor torque. Eigenvectors 10 and 11 associated with eigenvalues 10 and 11 represent the transverse torque needed to compensate for the center of mass not being directly on the rotor shaft axis and the forces and torque generated by aircraft forward motion and resulting air drag, respectively. It is understood that the values of the magnitude and orientation of the eigenvectors in this example are specific to the S-76® aircraft. However, the analytical process can be applied to other rotary wing aircraft. 
     Once the operating conditions are determined in steps (1)-(3) above, the actuator placement and orientation scheme is chosen. The actuator configuration should satisfy the eigenvector forces and orientation for the eigenvectors associated with eigenvalues 10-12. A system design for the S-76® is shown schematically in FIGS. 7-9. Each of the four corners of the transmission mounting footprint  60  represents a gearbox mounting foot  30 . The system utilizes a set of two, single acting actuators  36  per mounting foot  30  located in the predominantly vertical and predominantly horizontal directions. The vertical actuators  36  support the vertical lift load and pitching moment and the horizontal actuators  36  support the load resulting from torsion around the rotor shaft. As shown in FIGS. 7-9, actuator  36  mounting orientations off-vertical and off-horizontal were used to cover all of the eigenvectors of interest to support the smaller secondary static loads in the other three degrees of freedom. The vertical actuators  36  were angled forwardly and inwardly so that the principal axis of the actuators  36  formed an angle of about 10° and 5°, respectively, with the vertical axis. The right, aft horizontal actuator  36  was downwardly and outwardly angled so that the principal axis of the actuator  36  formed an angle of about 3.7° and 5°, respectively, with the central horizontal axis. The principal axes of the other horizontal actuators  36  are angled at about 15° relative to the central horizontal axis of the aircraft (FIG.  9 ), the left aft and right forward actuators inwardly and the left forward actuator  36  outwardly. 
     An iteration of Step (4) is performed until actuator size and orientation satisfy the loads under all operating conditions and subject to any physical mounting constraints for the particular application. The end result of the process defined above was an ATM actuator mounting scheme which defined the placement of two actuators at each transmission foot, the force requirements for the actuators and the detailed actuator mounting orientations. Dynamic hub loads for the S-76® design were measured using sensors installed on the S-76® rotor hub. These sensors measured the dynamic loads experienced by the transmission during flight. The magnitude of the measured dynamic loads was less than 10 percent of the largest static loads. The dominance of the static loads justifies the use of the static analysis as the principal design methodology. The magnitude of the dynamic loads was added to magnitude of the static loads described above to provide the overall actuator sizing. The resulting selection was tested by calculating the actuator forces required for each of the 29 flight cases. Small adjustments to actuator orientation and size were made until all 29 flight cases could be individually supported by the actuator system. 
     While the results of the system analysis according to the present invention has been presented with respect to the S-76® helicopter, it is understood that the system analysis is applicable to other helicopters and environments. Each make and model of helicopter to which this active vibration and noise reduction system of the present invention is applied will have a different set of actuator placement, orientation and size requirements due to that helicopter&#39;s unique design and flight envelope. The analysis, along with the engineering constraints considered for each make and model, will yield the appropriate actuator placement, orientation and size. 
     A preferred arrangement of a portion of the ATM system  34  is schematically shown in more detail in FIG.  10 . Each of the mounting feet  30  has an ATM  32  containing two ATM actuators  36  for canceling low frequency vibrations and for controlling the position of the foot  30  relative to the airframe. Six valves  48  are provided for controlling the eight ATM actuators  36 . The position of the two forward vertical ATM actuators  36  is preferably controlled together from a single hydraulic control valve  48 , Valve  1 . The two aft vertical ATM actuators  36  are individually controlled by dedicated hydraulic control valves  48 , Valve  2  and Valve  3 . The two forward horizontal ATM actuators  36  are also controlled together from a single hydraulic control valve  48 , Valve  4 . The two aft ATM actuators  36  are individually controlled by dedicated hydraulic control valves  48 , Valve  5  and Valve  6 . The six valves  48  provide fluid flow to the eight actuators  36  to control the position of the gearbox  22  relative to the helicopter&#39;s airframe in the three linear directions and about the three axes of rotation. Alternatively, each ATM actuator  36  may be associated with its own dedicated valve  48  so that the controller  44  and the actuation system  46  can vary the operational characteristics of each actuator individually. However, such individual actuator control has been found to be unnecessary and may be counterproductive. With eight control valves for the eight actuators, it is possible for the actuators to oppose one another resulting in excessive control and inefficient use of the hydraulic system  46  provided by the ATM system  34 . 
     FIGS. 11 and 12 illustrate the attachment of a gearbox mounting foot  30  to the airframe  28  using the ATM  32 . As shown schematically in FIG. 11, the mounting foot  30  is connected to the ATM  32  and the ATM is connected to the airframe  28 . The ATM  32  comprises an isolation frame  62  and a foot attachment  64 . The isolation frame  62  is fixed to the helicopter&#39;s airframe  28  and surrounds the foot attachment  64 . The gearbox mounting foot  30  is connected directly to the foot attachment  64  which is suspended from the isolation frame  62  by two ATM actuators  36  mounted on the isolation frame. One ATM actuator  36  is shown positioned substantially vertically on top of the isolation frame  62  and controls the vertical movement of the foot attachment  64 . However, depending on the mounting configuration, the vertical ATM actuator  36  could also be located below the mounting foot  30 . The other ATM actuator  36  is positioned substantially horizontally on a side of the isolation frame  62  and substantially parallel to the longitudinal axis of the helicopter for controlling the horizontal movement of the foot attachment  64 . FIG. 11 shows how the ATM  32  structurally isolates the foot attachment  64 , and thus the gearbox foot  30 , from the airframe  28 . FIG. 12 shows a detailed, exploded view of the ATM  32 . 
     A preferred arrangement of the ATM  32  with associated hydraulic valves  48  according to the present invention is shown in more detail in FIGS. 13 and 14 relative to one of the aft mounting feet  30 . Each ATM actuator  36  has a dedicated valve  48 . The passive noise isolator  58  is depicted as a tuned stub assembly, described more fully below. 
     As discussed above, hydraulic fluid flow to the ATM actuator  36  is actively oscillated to reduce the vibration transmitted from the gearbox feet  30  to the airframe. However, the quasi-static flight loads and displacements must be transmitted to the airframe. The applied quasi-static flight loads can change in magnitude and direction at frequencies up to 2 Hz depending on pilot inputs into the aircraft. Hence, the actuator  36  must be designed to accommodate the transmission of these large quasi-static flight loads. The magnitude of these loads can be quite high, approximately representing the aircraft weight amplified by the maneuver the aircraft is undergoing. For example, in a Sikorsky Aircraft Corporation S-76® aircraft, the quasi-static flight loads have a magnitude of about 8000 lbs. on each actuator  36 . The ATM actuator  36  must be designed to accommodate such loads while limiting gearbox motions to only about ±0.050 inches about a static position in order to avoid excessive misalignment of the engine transmission shaft. The ATM actuator  36  according to the present invention is also designed to prevent transmission to the airframe of small vibratory loads, e.g., 500 lbs. between about 16 Hz and about 50 Hz. These are the vibratory loads which cause the vibrations that are the most bothersome to the passengers and crew within the aircraft. 
     The actuator  36  is also designed to passively isolate the vibratory and acoustic loads which are applied transverse to the central longitudinal, or “active”, axis of the actuator. This is achieved by designing the actuator  36  so that the transverse stiffness of the actuator is low through the use of elastomeric bearing elements, discussed in greater detail herein below. Due to the configuration of the two actuators  36  at each foot  30 , each actuator&#39;s transverse axis is the other actuator&#39;s active axis. Thus, the actuator&#39;s  36  transverse axis may be passive because the other actuator  36  accommodates a particular load along that axis. 
     A preferred hydraulic ATM actuator  36  for use in accordance with the present invention is shown in FIG.  15 . The actuator  36  includes a housing  66  with a mounting flange  68  for attaching the actuator  36  to the airframe or a support structure. The actuator  36  also includes a mounting member  70 , depicted as a threaded stud, that is configured to attach to the gearbox foot attachment  64 . It is understood that other methods of attachment may be used with the present actuator design. 
     Referring to FIGS. 16 and 17, the mounting member  70  is attached to a piston  72  which is adapted to slide within a sleeve  74  which is secured within a recess in the housing  66 . The actuator  36  is preferably a lap-fit type actuator. Lap-fit actuators are well known in the art. Lap-fit actuators do not include any sliding seals. Instead, the lap-fit actuator is designed with extremely close tolerances (i.e., millionths of an inch) between the piston and the outer sleeve. As such, there is a close sliding engagement between the outer surface of the piston  72  and the inner surface of the sleeve  74 . Those skilled in the art of actuators are well aware of the design specifics of a lap-fit type piston arrangement and, therefore, no further details are needed. Suitable lap-fit piston and sleeve assemblies are available from Moog Inc., of East Aurora, N.Y., U.S.A. 
     Since manufacturing a piston  72  to meet the tolerances needed for a lap-fit is difficult, the outer surface of the piston in the present embodiment is formed with two raised annular surfaces  76  which provide the close sliding engagement between the piston and the sleeve  74 . A static seal, such as an O-ring  78 , is located between the sleeve  74  and the housing  66  to prevent hydraulic fluid from passing therebetween. 
     As described above, the actuator  36  undergoes both axial and transverse loads and motions. The loads and motions applied by the transmission foot  30  to the actuator  36  through the stud  70  comprise an axial quasi-static load and vibratory motion along “A”, and transverse vibratory motion along “T”. As described above, the quasi-steady axial load on a typical S-76® aircraft manufactured by Sikorsky Aircraft Corporation is approximately 8,000 pounds. The vibratory motion along “A” and “T” is approximately ±0.050 inches which stems from allowing the transmission to move in its own inertial frame. The transverse vibratory motion produces a transverse vibratory load transmitted to the airframe whose magnitude is governed by the inherent stiffness of the elastomeric bearing assembly. In the S-76® aircraft, this load is about 75 pounds. The combined axial loads and transverse motions result in a shear loading and a moment loading on the actuator  36 . The shear and moment loads on the actuator  36  cause transverse loads and cocking moments on the lap-fit piston  72 . Lap-fit type pistons  72  are not intended to see these types of loads since such loads can cause the piston to bind up and wear quite rapidly. Thus, a lap-fit piston could not be used in a conventional actuator  36  to react these applied loads. However, the present invention utilizes a novel bearing assembly  80  which is designed to react the shear and moment loads applied to the actuator  36  by the gearbox foot  30  and prevent the loads from being fully imparted to the lap-fit piston  72 . 
     As best seen in FIGS. 16 and 19, the bearing assembly  80  is located on top of the sleeve  74  and generally comprises an outer bearing member  82  which partially houses or supports three resilient bearings: a thrust bearing  84 , a radial journal bearing  86 , and a spherical bearing  88 . The outer bearing member  82  is partially housed within the housing  66 . The outer bearing member  82  is attached to the housing  66  using any conventional method. In the illustrated embodiment, the outer bearing member  82  has an integral flange  90  which contacts the upper end surface of the housing  66 . A clamping ring  92  fits over the flange  90  and is attached to the housing  66  with conventional fasteners (not shown) thereby securing the bearing assembly  80  and sleeve  74  in the housing  66 . The flange  90  presses against an O-ring  94  seated in an annular groove in the top of the housing  66 . In one embodiment of the invention, the housing  66  and clamping ring  92  are made from stainless steel material, although aluminum is preferable. The stud  72  is preferably made from aluminum material. The piston  72  and sleeve  74  are preferably made from stainless steel. 
     The thrust bearing  84  is mounted below the stud  70  and is preferably a substantially planar elastomeric bearing. Elastomeric bearings are well known in the art and generally comprise alternating layers of elastomer and nonresilient metal shims (not shown). The number of elastomer layers and shims is determined by the applied loads and desired flexibility of the elastomeric bearing. A suitable bearing assembly for use in the present invention is manufactured by Lord Corporation of Cary, N.C., U.S.A. The thrust bearing  84  is designed to be stiff, or rigid, in the axial direction and soft, or flexible, in the lateral or transverse direction. In one preferred embodiment, the thrust bearing has an axial stiffness greater than about 300,000 lbf/in, and a transverse stiffness less than about 1,000 lbf/in. The thrust bearing  84  is preferably located between and attached to the inner end of the stud  70  and a thrust mounting plate  98 . The attachment of the bearing  84  to the stud  70  and the mounting plate  98  is through any conventional means, such as an elastomer to metal adhesive. 
     The thrust mounting plate  98  is attached to or formed on the upper end of an inner bearing member  100 . At least a portion of the lower end of the inner bearing member  100  is located within the outer bearing member  82 . The radial journal bearing  86  is located between the inner bearing member  86  and the outer bearing member  82  and is preferably a substantially cylindrical elastomeric bearing. The radial journal bearing  86  is preferably stiff in the radial direction and soft axially. In one preferred embodiment, the radial journal bearing  86  has a radial stiffness greater than about 300,000 lbf/in, and an axial stiffness less than about 3,000 lbf/in. 
     The spherical bearing  88  is located within and engages with an inner spherical surface of the inner bearing member  100 . The spherical bearing  88  fits over and is attached to a spherical bearing support  102 . The spherical bearing  88  is preferably an elastomeric bearing which is stiff axially and soft rotationally. In one preferred embodiment, the spherical bearing has an axial stiffness greater than about 500,000 lbf/in, and a rotational stiffness less than about 15 in-lbf/deg. The spherical bearing support  102  is engaged with the piston  72  through a piston rod  104 , or stinger. The piston rod  104  is preferably made from a metallic material, such as steel, and is attached to the spherical bearing support  102  and piston  72  through any conventional means known to those skilled in the art, such as a snap ring  105  and bolted connection  106 , respectively. 
     When the bearing assembly  80  is assembled, the centers of rotation of the spherical bearing  88  and the radial journal bearing  86  are as close as possible (e.g., &lt;0.03 inches). This coincidence of centers of rotation is required to ensure stability of the elastomeric bearing assembly  80  and avoid buckling of the ATM actuator  36  under axial load. Additionally, the centers of rotation of all three elastomeric bearings  84 ,  86 ,  88  are as nearly co-axial as possible (e.g., &lt;0.03 inches). Further, the co-axial centers of rotation are substantially coincident along the longitudinal axis of the piston  72 . In this way, offsets in centers of rotation do not combine with the high quasi-static axial load to produce large transverse loads on the piston  72  and sleeve  74 . 
     The preferred bearing assembly  80  provides a novel mechanism for reacting to and isolating transverse motions (e.g. vibrations and noise) stemming from the gearbox foot  30  from creating transverse loading on the piston  72  and sleeve  74 . As stated above, the applied transverse motions result in shear and moment loads on the stud  70  which can then ripple through a bearing assembly. Additionally, transverse motions together with the high quasi-static loads along the longitudinal axis of the ATM actuator  36  further increase moment loads on the stud  70 . For a properly designed bearing assembly  80 , shear loads caused by transverse motions are low due to the low transverse stiffness of the thrust bearing  84 . The low shear loads are then further prevented from coupling into the piston  72  and sleeve  74  due to the high transverse stiffness of the radial journal bearing  86  which protects the inner bearing member  100  from radial motion with respect to the outer bearing member  82 . Moments placed on the stud  70  due to the shear and off center quasi-static loads are prevented from reaching the piston  72  and sleeve  74  by the spherical bearing  88 , which is very soft rotationally. As a consequence, only small moments are transferred through the spherical bearing  88  to the spherical bearing support  102  and, thus, to the piston rod  104 . The lower end of the piston rod  104  reacts to these small moments as a small transverse or side load between the piston  72  and sleeve  74 . Further, the design of the bearing assembly  80  is such that axial loads are transmitted through the thrust bearing  84  and spherical bearing  88  directly to the piston  72 . The cumulative axial stiffness of the thrust bearing  84  and the spherical bearing  88  must be as high as possible, preferably, greater than 100,000 lbf/in, to preclude excessive compression of these members under quasi-static load. Excessive compression would cause unwanted flow demands upon the hydraulic valve  48  and the supply  54 . 
     Pressurized hydraulic fluid is channeled through an inlet  108  to a chamber  110  and against the head of the piston  72 . Since a lap-fit piston does not include seals, pressurization is obtained by a close-fit sliding interface between the piston  72  and the sleeve  74 . However, even with this close fit, hydraulic fluid will escape between the piston  72  and the sleeve  74  as the piston  72  moves relative to the sleeve  74 . Hydraulic fluid can damage the elastomeric material of the bearing assembly  80 . 
     According to the present invention, a diaphragm  112  is provided for physically separating the bearing assembly  80  from the sleeve  74  and the piston  72  and substantially preventing slightly pressurized hydraulic fluid which leaks past the piston  72  from contacting the bearing assembly  80 . As seen in FIG. 17, the diaphragm  112  is flat and ring-shaped, and is made from a material that is benign to the hydraulic fluid used. Preferably, the diaphragm  112  is made from an elastomeric material. For example, nitral rubber is suitable for use with Mil-H-5606 fluid. 
     The inner diameter of the diaphragm  112  (FIG. 18) is attached to the spherical bearing support  102  and the outer diameter is attached to the outer bearing member  82 . Two annular sealing rings  114 ,  116  are bonded with urethane to the inner and outer edges of the diaphragm  112  on the piston side. Additional urethane and peripherally-arranged machine screws  118  through the two sealing rings  114 ,  116  bond the inner and outer edges of the diaphragm  112  to the spherical bearing support  102  and the outer bearing member  82 , respectively. Thus, a tight seal is formed between the diaphragm  112  and the spherical bearing support  102  and between the diaphragm  112  and the outer bearing member  82 . This tight seal prevents migration of hydraulic fluid from the piston side of the diaphragm  112  to the bearing side of the diaphragm. As the piston  72  moves, the diaphragm  112  accommodates slight stretching without undue stress. 
     The space  120  between the diaphragm  112  and components of the bearing assembly  80  are preferably filled with polybutene. Polybutene is preferred because it is benign to natural rubber, is not petroleum based, never hardens, is non-toxic, and does not react unfavorably with the elastomer used in the bearing assembly  80 . The polybutene also prevents excessive deformation of the diaphragm  112  when subjected to the low pressure caused by leakage of hydraulic fluid past the piston  72 . Excessive deformation of the diaphragm  112  would risk excess stress along the sealed edge surfaces of the diaphragm  112 . 
     The present invention also accounts for leakage of hydraulic fluid past the piston  72 . As seen in FIG. 18, the upper end of the sleeve  74  has an outward flange having an annular groove  124  spaced from the periphery of the flange. A plurality of axial openings  126  in the groove  124  pass through the flange portion of the sleeve  74 . The openings  126  communicate with a corresponding annular groove  128  in the housing  66 . The housing  66  also has a plurality of axial openings  130  each leading to a radial passage  132 . The passages  132  opens into a circumferential groove  134  in the outer periphery of the housing  66 . 
     A fluid outlet passage  136  in the mounting flange  68  communicates with the groove  134 . The fluid outlet passage  136  receives a hydraulic fitting  138  which connects to a drain line (not shown) to remove the leaked fluid. Two additional outer circumferential grooves  140  in the housing  66  flank the fluid receiving groove  134 . Sealing O-rings  142  are located in each of the two outer circumferential grooves  140  to prevent hydraulic fluid from leaking between the housing  66  and mounting flange  68 . Therefore, leaked hydraulic fluid has a free flow path from the sleeve  74  to the periphery of the housing  66 . Moreover, because the hydraulic fitting  138  located on the mounting flange  68  is always in fluid communication with the central circumferential groove  134  regardless of the relative rotational positions of the sleeve  74  and the mounting flange  68 , the hydraulic fitting  138  is always properly placed with respect to the preformed and pre-fitted drain line. There is no rotational positioning requirement for the sleeve  74 . 
     During assembly of the actuator  36 , it is important that the axis of the sleeve  74  be concentrically located with respect to the central axis of the bearing assembly  80  so that the axial quasi-static load applied to the stud  70  is directed along the central axis of the sleeve  74 . Therefore, the location of the bearing assembly  80  with respect to the axis of the sleeve  74  is a critical assembly parameter. 
     To ensure alignment of the bearing assembly  80  and the sleeve  74  an intermediate reference surface is needed. The reference surface  147  is an internal diameter of the housing  66 . During manufacture of the sleeve  72  the concentricity of the sleeve bore which accepts the piston  72  is held to within about 0.0005 inches of the sleeve&#39;s outside diameter which in is contact with the reference surface  147 . Additionally, the outside diameter of the sleeve is held to within 0.003 inches of the diameter of the reference surface  147 . During manufacturing of the outer bearing member  82 , the outside diameter of the outer bearing member is also held to within 0.003 inches of the diameter of the reference surface  147 . During manufacturing of the bearing assembly  80 , the total runout of the inner bearing member  86 , spherical bearing support  102 , and stud  70  is held to within 0.002 inches relative to the outer diameter of the outer bearing member  82 . Adherence to these tolerance will ensure that stability of the ATM actuator is maintained and side loads between the sleeve  74  and piston  72  remain low. 
     The passive isolator  58  comprising the tuned stub assembly is shown in FIG.  20 . The tuned stub assembly comprises a manifold  152  and two tubes  154  connected to the manifold. One end of each tube  154  opens into the manifold  152  and the other end of the tube is closed using a compression cap  156 . The hydraulic line  56  from the valve  48  enters one side of the manifold  152  and a hydraulic line  158  exits the manifold  152  for carrying fluid to the ATM actuator  36 . The connections all comprise compression fittings, although any fluid-tight connection is suitable. The tubes  154  are preferably a stiff material such as steel, which is otherwise ordinarily used in hard plumbed hydraulic systems. Although two tubes  154 , or “stubs”, are shown, the number of stubs  154  used depends on the number of tonal frequencies desired to be reduced. 
     The tuned stubs  154  operate on the well known principle of standing compressional waves within a tube closed at one end. The tuned system has a very low impedance to small movements of fluid at the mouth of the tube that move at the natural frequencies of the tuned stub. Therefore, when tuned to vibratory pressure waves at a particular frequency, the stiffness of the fluid at the mouth of the tube appears to be low, which translates into an apparently low stiffness at the associated hydraulic actuator. In a helicopter application, the stubs are tuned to reduce the “gear clashing” fundamental frequency and its odd integer tonal frequencies. The number of teeth on the gears and the speed of the gears determine the gear clashing frequencies. This frequency is high compared to that of vibrations, typically greater than 500 Hz, although somewhat lower frequencies are occasionally experienced. 
     The fundamental resonant frequency of the stub to which the stub is “tuned” is a function of the speed of sound in the hydraulic fluid and the tube&#39;s length and is calculated as follows:        f   =     c     4      L                              
     where f=target frequency (Hz) 
     c=the speed of sound in the hydraulic fluid (m/s) 
     L=fluid path length from the entrance of the fluid into the actuator to the inside surface of the tuned stub cap (m) 
     Since the target frequency and the speed of sound in the fluid are known, the fluid path length is calculated to determine the length of the tube. For example, assuming a stiffness null is needed at 1 kHz and the speed of sound in the hydraulic fluid is 1450 m/s, solving the equation above for L yields a fluid path length from the actuator to the cap of 0.363 m (14.3″). 
     The addition in the fluid column of elbows, constricted areas, varying diameters, and the like, all lead to complex hydraulic dynamics which worsen as frequency increases. Therefore, it is preferable that the piping length from the actuator to the tuned stub be as short as possible to minimize the complexity of the hydraulic dynamics and ensure the hydraulics at the actuator is soft at the tonal frequencies. This “softness” is based on the ability for the fluid in the chamber  110  to communicate with the fluid at the mouth of the tuned stub  154  in the manifold  152 . Therefore, the stub  154  is preferably as close to the hydraulic fluid inlet  108  as possible so there is no interaction due to elbows, fluid line diameter reductions, and the like, which introduce, in addition to complexity, damping effects reducing the stiffness null and thereby decreasing the effectiveness of the tuned stub. 
     During initial design, the tube length can also be accurately determined using a compression cap  156  with a movable piston  158  having a diameter equal to that of the inside diameter of the tube. FIG. 21 illustrates the compression cap  156  with a threaded central hole for receiving a threaded rod  160  attached to the top of the movable piston  158 . The threaded rod  160  extends from the movable piston  158  through the compression cap  156  and is accessible from outside of the tube  154  for moving the piston in the tube and effectively adjusting the fluid column length (L). Once the appropriate fluid length is found creating a null in the stiffness at the desired frequency, the rod  160  and piston  158  are removed and the tube is cut to length so that the inside surface of the compression cap  156  is at the same position as the piston  158 . 
     The preferred inside diameter of the tuned stub  154  relates to the diameter of the piston  72  of the actuator  36 . This is because the stiffness of the tuned stub as seen by the actuator scales as the ratio of the piston  72  cross-sectional area to the cross-sectional area of the tuned stub  154  squared:          K   actuator     =         (       A   piston       A   stub       )     2          K   stub                              
     Practically, one can expect to achieve a 10:1 reduction (˜20 dB) of the entrained fluid static stiffness using a tuned stub  154 . The reduction is primarily limited by damping which results from subtle causes such as burrs on tubes and fittings, exposed internal threads, slight changes in fluid column diameter, and the like, that lead to turbulence in the line. Thus:            K   dynamic       K   static       =     1   10                            
     Since K dynamic  is the dynamic stiffness of the stub at the tuned frequency and K static  is the static stiffness of the stub,          K   actuator     =         (       A   piston       A   stub       )     2            K   static     10                              
     K static  is calculated as follows:          K   static     =       ρ                   c   2          A   stub       L                            
     where ρ=fluid density (kg/m 3 ) 
     c=the speed of sound in the hydraulic fluid (m/s) 
     L=fluid path length from the actuator to the tuned stub cap (m) 
     A stub =cross-sectional area of stub (m 2 ) 
     Therefore,          K   actuator     =         A   piston   2       A   stub              ρ                   c   2         10      L                                
     Solving for the cross-sectional area of the stub:          A   stub     =         A   piston   2       K   actuator              ρ                   c   2         10      L                                
     leads to the stub&#39;s inner diameter for optimal noise reduction. 
     The tuned stubs isolate noise due to high frequency transmission gear clashing by providing low mechanical stiffness at selected frequencies where high levels of noise occur. The value used for K actuator  is unique to each make and model of helicopter. In general terms, one needs to determine the airframe stiffness at the frequency of interest as seen by the gearbox foot  30  along the active axis of the ATM actuator  36 , K actuator  is then set at 1/10. For the ATM system for the S-76® helicopter, K actuator  was &lt;60,000 lbf/in. 
     Overall, the ATM system acts as a filter, creating low stiffness regions in the transmission&#39;s mounting stiffness at selected frequencies of interest. The effect is to decouple the shaking transmission at those frequencies from the helicopter&#39;s airframe. At frequencies of the vibration and noise generated within the transmission itself, the actuator system makes the stiffness seen by the transmission feet, looking into the actuators, to appear to be very low by both active and passive means. The ATM system actively creates low stiffness notches at the rotor induced vibration frequencies and passively creates similar low stiffness notches at the gear-clashing induced noise frequencies. 
     The present invention provides a novel actuator and system for reducing vibrational and noise transmission from a helicopter gearbox footing to the airframe. The ATM system was tested on the gearbox footing of a Sikorsky Aircraft Corporation S-76® aircraft. Reductions of about 13 dB in cabin noise and about 25 dB in vibrations have been achieved using the present invention. While the invention is illustrated and described as being used for reducing vibration and noise transmission from a gearbox foot  30  to an airframe, it is not limited to that embodiment. On the contrary, the present invention can be used to address vibration and noise transmissions from a variety of other components in various types of machines and aircraft. Although the invention has been described and illustrated with respect to the exemplary embodiments thereof, it should be understood by those skilled in the art that the foregoing and various other changes, omissions and additions may be made therein and thereto, without parting from the spirit and scope of the present invention. For example, a Hemholtz resonator could be used in place of the tuned stubs although this approach would address only the fundamental and not the tonal frequencies.