EP0704037A4

EP0704037A4 - Inertial coupling mount

Info

Publication number: EP0704037A4
Application number: EP94917347A
Authority: EP
Inventors: Scott K Miller; Andrew J Langley
Original assignee: Noise Cancellation Technologies Inc
Current assignee: Noise Cancellation Technologies Inc
Priority date: 1993-05-13
Filing date: 1994-05-11
Publication date: 1996-09-25
Also published as: WO1994027083A1; JPH08508815A; EP0704037A1

Abstract

An inertial coupling mount with active vibration isolation means between the source of vibration and the base to which it is mounted is disclosed. Inertial coupling (30) in series with actuator (4) provides vibration control force to effect high-frequency, engine-induced, vibration reduction. Coupling (30) reduces displacements to which actuator (4) is exposed, enabling effective control throughout very large, road-induced mount displacements, while at the same time protecting the actuator.

Description

INERTIAL COUPLING MOUNT

This invention relates to an inertially-coupled active engine mount. The object of the invention is to effect vibration isolation between a source of vibration and the base to which it is mounted. Usually the source of vibration will be a vehicle engine and the base will be the frame or body of the vehicle. Design of traditional vibration isolation mounts, which is well-known, involves making a compromise between control of the engine's position and the isolation of vibration. The vibration across an automotive engine mount below a certain frequency (about 15 Hz) can generally be attributed to movement of the car body while vibration above this frequency is generally due to imbalance, reaction forces from the piston motion, or torque pulsation forces generated by the engine. Several authors have proposed that control forces be applied to various parts of a traditional mount to minimize the effect of this compromise. In other words, a fairly stiff mount, which effectively controls engine position, can be made to effectively isolate high frequency vibrations emanating from the engine in the same way an extremely soft mount would normally do. The proposed invention circumvents limitations of each of the prior art implementations in achieving this objective.

Background Art Three inventors, Idigkeit et al (U.S. Patent No. 4,638,983), Andra (U.S. Patent

No. 4,693,455) and Hoying et al (U.S. Patents 4,778,158 and 4,789,142) have proposed adding active control forces to a conventional hydraulic engine mount. In a conventional hydraulic engine mount, two fluid-filled chambers are separated by a passage through which fluid is communicated between the chambers. The chamber stiffnesses are chosen such that the fluid in the passage will have high velocity, and hence high viscous losses, at the "bo nce" frequency of the engine. In this way, good motion control is obtained and the mount stiffness is still fairly low at higher frequencies. A problem with this type of mount is that forces related to the fluid inertia increase with frequency, causing the mount stiffness to increase as well. The above authors have added actuators which cause pressure fluctuations in one of the fluid-filled chambers. These pulsations are chosen such that dynamic forces across the mount are minimized. Sometimes only the fluid- related forces are reduced and sometimes the net force across the mount is caused to be zero. In either case, a reduction in the dynamic force transmitted across the mount is obtained. Since the actuator is coupled across the mount through the resonant fluid cavity, this arrangement will be referred to as a resonant coupling arrangement.

Another inventor, Garnjost (U.S. Patent No. 5,067,684), proposes a mount unlike the hydraulic engine mounts in which an actuator and damper are interposed across the simple rubber engine mount. The"active" force across the mount is applied through the damper. Garnjost notes that the damper, when properly sized, causes the actuator displacements to be reduced at low frequencies. Since the actuator is coupled across the mount through a damper, this arrangement will be referred to as a viscous coupling arrangement. Accordingly, it is an object of this invention to provide vibration isolation between a source of vibration and the base to which it is mounted.

Another object of the invention is to reduce the complexity of the vibration control actuator by reducing the displacements to which it is exposed.

This and other objects will become apparent when reference is had to the accompanying drawings in which

Figure 1 shows, schematically, a mount configuration used by inventors Idigkeit, Andra and Hoying et al.

Figure 2 shows the configuration described by inventor Garnjost.

Figure 3 shows the preferred embodiment of the present invention. Figure 4 shows a cutaway view of a component of the present invention, a piston and cylinder as an inertial coupling.

Figure 5 is a diagram showing a plot of the ratio of engine displacement to actuator displacement versus frequency.

Figure 6 shows a cutaway view of an embodiment in which the shear coupling is a rod.

Figure 7 shows a cutaway view of an alternative version of the embodiment of Figure 6.

Figure 8 shows a cutaway view of the present invention in use on a conventional hydro-mount. Referring to the drawings, Figure 1 shows the shortcomings of the approach suggested by Idigkeit, Andra and Hoying et al. It shows the arrangement schematically, replacing physical elements with their dynamic equivalents. Spring 1 and damper 2 are the elastomeric stiffness and damping of the mount. Spring 3 represents the "bulge" stiffness of the mount. This element represents the pressure forces exerted by the fluid in the mount as a result of displacements of fluid or the various fluid-displacing mount components. Actuator 4 is restrained to displace only a small amount by stops 5. The placement of the support 20 under beam 6 represents the area ratio between the fluid passage and the main chamber and mass 7 represents the mass of the fluid in the passage. Spring 8 and damper 9 represent the stiffness restraining the fluid in the passage and the viscous resistance to flow through the passage, respectively. Plate 10 represents the vehicle body while mass 11 represents the portion of the engine mass supported by that particular mount. It can be seen that when the mass 11 resonates, the force through spring 3, which acts to restrain the engine, is high. This occurs when the engine has been caused to move at large displacements relative to the frame by forces input from the road. This high force through spring 3 only occurs when the actuator has been "shorted" by the stops 5. Therefore, the actuator 4 is unable to apply control forces across the mount under conditions of high engine displacement. The design of Garnjost is shown schematically in Figure 2 showing engine mass

11, spring 1, damper 2 and actuator 4 acting on master piston and cylinder 16, and slave cylinder 15. Flow through restriction 14 results in energy dissipation, so a damper is effectively in parallel with the mount. Since the control frequency range is not substantially higher than the "bounce" frequency range, the actuator still needs to undergo substantial displacements. The viscosity of various potentially usable fluids may vary by a factor of 10 or more over the working temperature range of the device, causing significant variation in mount behavior and effectiveness. It is significantly advantageous to be able to use a short-stroke actuator, whether that actuator is hydraulically or electrodynamically operated. The object of the current invention is to reduce the requirement that a high displacement actuator be used and to provide a means of controlling high frequency vibrations even as the engine and body undergo large displacements. It is advantageous to use an actuator with low internal stiffness to minimize transmission of very high frequency vibration across the mount above the control bandwidth.

Description of the Invention

Figure 3 shows, schematically, how the invention is arranged. Many of the elements of Figure 1 are present and the actuator is moved into a position in parallel with the mount stiffness and damping and is placed in series with an inertial coupling 12 and a shear coupling 13. The inertial coupling is a device which transmits a force which is proportional to the acceleration of the device's two terminals. The shear coupling transmits force primarily along its axis while having very low stiffness in the plane normal to its axis. One embodiment of the inertial coupling is shown in Figure 4 in which the coupling 12 is a hydraulic cylinder 31 with a piston 32 having openings 33 which allow fluid to pass between its chambers. The openings are large enough that the fluid viscosity does not impede the flow significantly, but are small enough, compared to the piston area, that the acceleration of fluid through the openings causes a force to be transmitted across the coupling. The shear coupling can consist of a device as disclosed in Chaplin's U.S. Patent 4,600,863, or a stack of many pieces of rubber which have a high area to thickness ratio separated by stiff material, or a hydraulic device constructed of rubber and a fluid as used in common engine hydraulic mounts. It could also be as simple as a flexible rod. The advantage to using an inertial coupling rather than the viscous coupling proposed by Garnjost is that the force transmitted through the device below the "cutoff' frequency, which is determined by the characteristics of the mount and the coupling, falls off much faster for an inertial device than a viscous device. Both these arrangements are superior to the resonant coupling arrangement from the point of view of actuator displacement minimization. As an example, consider Figure 5, which shows, conceptually, how the displacement of an actuator with the various types of coupling depends on displacements applied to the engine. It is clear that at low frequencies, at which the road- and driving- induced displacements are high, the inertial coupling protects the actuator from those severe displacements, hence allowing the mount to continue to control high frequency vibration. At high frequencies, at which control of engine-induced vibration is desired, the actuator displacements need only be small to apply sufficient force to cancel the vibration.

Three preferred embodiments of the invention are now discussed although others will be obvious to those skilled in the art. The first involves attaching the series assembly of actuator, inertial coupling and shear coupling directly in parallel to a simple rubber mount. This "active assembly" could either be internal or external to the rubber structure. In the embodiment of Figure 6, the shear coupling 40 is simply a long, thin rod which transmits very little force in directions perpendicular to its axis. The mount stiffness 41 is provided by molded rubber, which supports the engine's weight and reacts against its torques. The engine is attached at engine attachment means 42 and the bottom of the mount rests on the vehicle body 10. The actuator 4 drives across the mount through the inertial coupling 12, the lower end of which is attached to the bottom of the mount and the car body 10. In the Figure 7 embodiment, the shear coupling action is provided by the fact that the clearance between the plunger 51 and rigid sides 52 of the mount is large enough for fluid 53 to flow unimpeded by viscous resistance, but small enough to induce high accelerations in the fluid. This enables rotation and off-axis motion of the piston while allowing unrestricted axial motion and "inertial" resistance to that motion. Figure 8 shows how a conventional hydro-mount might be modified to include the current invention without sacrificing the beneficial behavior of the hydro-mount at low frequencies. The hydro-mount might be modifed to include the current invention without sacrificing the beneficial behaviour of the hydro-mount at low frequencies. The hydro-mount contains two fluid cavities 62 and 63 which are connected by an orifice 66. The lower cavity 63 is bounded by a rubber diaphragm 65 on one side. In this embodiment, tube 61 is a device like that disclosed in the referenced Chaplin patent above except with low axial stiffness and high circumferential stiffness, and hence resistance to radial expansion. This "tube" would exert forces across the hydro-mount 60 as a result of motion of the actuator 4 and inertial coupling 12 without otherwise influencing the pressure within the mount's operating fluid cavities 62, 63 and 64. The pressure between cavities 62 and 64 could be equalized with small holes allowing communication between the cavities. The tube 61 could be constructed of a rubber tube reinforced with steel rings, a wire mesh, or other means such that it would contain fluid pressure without expanding radially, but would easily expand axially.

Having described the invention it will be obvious to those of ordinary skill in the art to make modifications and changes without departing from the scope of the appended claims.

Claims

1. An inertial coupling mount system, said system comprising: inertial coupling means, shear coupling means, force coupling means in series with said inertial coupling means and said shear coupling means, stiffness and damping means, said force coupling, inertial coupling and shear coupling means being mounted in parallel with said stiffness and damping means to thereby effect vibration isolation.

2. A system as in claim 1 in which actuator motion is minimized rapidly below a cutoff frequency which is higher than the mounted natural frequency of the equipment.

3. A system as in claim 2 in which said shear coupling means and said inertial coupling means are integrated.

4. A system as in claim 2 and including a small stroke actuator means adapted to only provide small displacements.

5. A system as in claim 2 wherein said inertial coupling, shear coupling and force coupling means are integrated in a hydro mount design.

6. A system as in claim 2 including means adapted to control throughout large, low frequency engine displacements.

7. A system as in claim 6 including an actuator means, said actuator means having a low internal stiffness to reduce the transmission of vibration at high frequencies outside a predetermined control bandwidth.

8. A system as in claim 2 wherein said stiffness and damping means comprises, molded rubber component means.

9. A system as in claim 2 and including one or more orifice means, whereby the force proportional to acceleration of the two ends is caused by accelerating fluid through one or more of said orifices. 10. A system as in claim 9 and including a cylinder means and piston means said piston means in motion in said cylinder means causing acceleration of fluid through orifices in said piston or between said piston and cylinder means.