Suspension and Vibration Isolation System for a Linear Reciprocating Machine

The present invention relates to vibration isolation systems and in particular to a dynamic vibration isolating suspension system which may be incorporated into a mass flow system of a linear reciprocating machine.

The term vibration refers to a continuing periodic motion of one or more particles or one or more rigid bodies with respect to a specified central reference or static equilibrium position. Rigid body mechanical vibration more particularly refers to the periodic motion of a body having mass which is supported by one or more elastic elements having stiffness (or conversely, resiliency).

Elastic element, as used herein, defines a device which imposes a force on the body tending to restore the body to its static equilibrium position if the body becomes displaced from its equilibrium position. Mass or inertia implies that the velocity imparted to the body during restoration of the static equilibrium position will cause the body to overshoot its equilibrium position. Mechanical vibration is the consequence of the interplay of mass and stiffness resulting in periodic displacement of the mass about its original static equilibrium configuration at the instigation of an exciting force.

The vibration of a solid body in contact with a fluid medium such as air, water, etc., imparts vibrations in the medium. When the rate of vibration of the solid body ranges between 30 and 15,000 hertz, the imparted vibration transmits energy in the form of an acoustic wave, audible to the human ear. As a practical consequence, mechanical vibrations produce undesirable noise, motion and forces.

Vibration isolation generally refers to the means used to reduce the transmittance of vibration of one body to another, or to reduce the transmittance of periodic (dynamic) forces from one body to another. The simplest form of a vibration isolator is a mechanical element or device having low stiffness (or high resilience) and negligible mass.

Examples of above-described isolators include springs and rubber blocks. This basic type of vibration isolator has been employed in suspension systems for linear reciprocating machines such as may be found in U.S. Pat. Nos. 3,666,219 and 2,894,678. A mass-elastic diagram generally representing this basic approach to vibration isolation is depicted in FIG. 1A. The transmissibility of such isolation systems is a function of the dynamic stiffness of the isolation device, the inertial mass of the machine or body being supported by the isolation device, and the frequency of the vibratory motion or periodic forces imposed on the inertial mass. The shortcomings of these passive isolation devices include amplified vibration at resonance frequency and high static deflections if acceptably low vibration transmission characteristics are to be achieved at normal operating frequencies. The term transmission here relates specifically to the transmitting of periodic motion or force from one solid body to another. Transmissibility is generally defined as the ratio of the force transmitted by the isolator means to the force applied by the exciting source.

Low transmissibility herein refers to a transmissibility value less than unity which is the desired objective of vibration isolation devices. Unfortunately, it has been substantially difficult to achieve low transmissibility with simple (passive) isolation systems while at the same time providing sufficient static stiffness for acceptable support means. Employed as a support system for a reciprocating machine, isolators of sufficient resiliency to satisfy the machine's dynamic isolation requirements frequently result in high static deflections and may lead to static equilibrium instability problems.

The problem of large (resonant) vibratory displacements at the resonant frequency may be alleviated by use of a damping system in parallel with the isolator such as a spring/dashpot combination as exemplified in U.S. Pat. No. 3,807,678. However, at frequencies above resonance, which is the normal operating frequency range of a properly designed isolation system the damping mechanism increases dynamic force transmissibility.

The present invention employs dynamic vibration isolation to achieve substantially reduced displacement and force transmissibility at typical operating frequencies while maintaining sufficient static stiffness (or rigidity) to support relatively heavy reciprocating machines such as a reciprocating compressor, for example, without unduly large static deflections. A further advantage of the present invention is the ability to substantially reduce displacement and force transmissibility regardless of the vibratory displacements experienced by the supported mass as a consequence of exciting forces acting on the mass. In other words, the static support stiffness of the dynamic isolation system can be high and the vibratory displacements of the supported mass can be large while the resultant dynamic force transmitted to the isolator foundation, such as a floor, for example, will still be very low.

A two-mass vibration isolator is conceptually capable of producing a transmissibility substantially lower than that of a conventional isolator having similar stiffness characteristics. An example of a 2-mass isolator is depicted in FIG. 1B. The 2-mass isolator shown in FIG. 1B is commonly referred to as a vibration absorber isolator. The characteristics of such a 2-mass system differ from the aforementioned simple (passive) isolator inasmuch as the two-mass system has two resonant frequencies. Within a small band of frequencies between the two resonant frequencies and at frequencies above the highest resonance the transmissibility of the two-mass system is smaller than that of simple single-mass isolators of similar stiffness.

The mass elastic model of FIG. 2 illustrates, in its most simplistic form, one embodiment of the isolator of the present invention. In contrast to the prior art 2-mass isolator, the present invention can be designed to yield essentially zero transmissibility at a frequency above the two resonant frequencies of the isolator system while concurrently maintaining substantially reduced transmissibility over a broad range of frequencies above the two resonant frequencies.

In a further embodiment, the present invention achieves substantially reduced transmissibility within a band of frequencies between the two resonant frequencies. In contrast to the isolation systems known in the art which reduce transmissibility only as a consequence of reduced vibratory displacement, the present invention achieves reduced transmissibility virtually irrespective of the vibratory displacements of the body (machine) being isolated.

In U.S. Pat. Nos. 3,091,103 and 3,202,388 the latter being a divisional of the former disclosure, the patentee applies the principles of a two-mass system to provide vibration isolation of periodic motion. As disclosed therein, a fluid vibration isolator has a variable volume bellows which expands and contracts upon applying periodic (vibrational) force. The inertial reaction of a fluid, which flows into (and out of) the bellows, balances the force applied to the body to be isolated. These isolators are designed to transmit constant (i.e. static) force while suppressing periodic (vibratory) forces.

In contrast, the vibration isolator of the present invention utilizes a mass flow conduit system having a prepositioned counterbalance mass to provide high static stiffness and essentially zero transmissibility of vibratory forces at the operating frequency of the machine.

FIGS. 1A and 1B are schematic representations of prior art vibration isolation systems.

FIG. 2 is a schematic representation of the vibration isolation system of the present invention.

FIG. 3 is a cut-away view in elevation of a compressor employing one embodiment of the invention's syspension system.

FIG. 4 is a cut-away axial view of the compressor and suspension system of FIG. 3.

Briefly stated, the suspension system of the present invention utilizes a particularly adapted conduit system in conjunction with a counter balancing mass to provide a dynamic vibration isolation system for a reciprocating machine having periodic vibratory motion. In-phase and antiphase spring reaction forces are established by the assembly of conduits at operating frequency to produce an essentially zero net resultant force acting upon the points of mounting to an outer casing, foundation or other support means. A low transmissibility of periodic motion and dynamic forces is achieved virtually irrespective of the vibratory displacements of the body or machine being isolated.

The present invention provides vibration isolation means for essentially eliminating transmittance of the periodic motion (vibration) and dynamic forces inherent to the operation of the reciprocating machine to the foundation supporting said machine and its casing. The result is a virtual elimination of vibration and an attendant reduction of noise transmitted outside the casing.

As illustrative thereof, in FIGS. 3 and 4, there is shown a linear reciprocating motor driving a compressor having inlet and discharge conduit comprising a part of its mass or fluid flow system. The assembly comprises a metallic outer casing 10 having a generally cylindrical central portion 12 and at least two generally hemispherical end sections 14 engaging said cylindrical portion 12 at 16. Said engaging may be accomplished by conventional means of rigidly affixing end sections 14 to the central portion 12 such as welding, riveting or the like. Though not illustrated, at least one end section may alternatively be removably affixed to central portion 12 by conventional means (not shown) such as nuts and bolts, threaded collars, or similar, to provide access to the interior of casing 10. A plurality of lower support members 18, affixed to underside of said casing 10, provide means for securing the compressor assembly to any suitable mounting support.

In one embodiment, a linear reciprocating compressor 20, which might be of the type illustrated in U.S. Pat. No. 3,937,600, is contained within said outer casing 10. Said compressor is described generally as having horizontally disposed symmetric compression chambers. A similarly situated symmetric piston has a principal direction of motion along a horizontal axis. The compressor may be driven by alternate linear reciprocating machines including, for example, linear electric motors, Stirling type engines or the like. Although compressor orientation and resultant vibration may be in directions other than the horizontal, the present invention's vibration isolator will be described in terms of motion along a horizontal axis. Motion in directions other than the horizontal is a logical extension of the forthcoming description and operation analysis.

A portion of the suspension system for compressor assembly 20, including linear reciprocating motor 20B, serves simultaneously as a mass flow conduit system in the operation of said assembly. In the embodiment illustrated in FIG. 3 a pair of resilient helical inlet tubes 22 are axially spiraled about the opposite end regions of compressor assembly 20. One end of each helical inlet tube engages the outer casing 10 through connecting means here shown as inlet tube collars 24, 24. Casing 10, here illustrated as at least partially enveloping compressor assembly 20, may equivalently comprise any structural support means for securing the suspension system thereto. Accordingly, casing 10 equivalently comprises any frame support means capable of supporting the linear reciprocating machine relative to ground. The inboard end of each of said helical inlet tubes 22 is rigidly affixed to the inlet port 28 of the compressor by means of an inlet collar 30 which is secured to said compressor assembly by means of a plurality of circumferential welds 32. In the embodiment of FIG. 3, non-rigid vertical suspension is provided by vertical support means 21. Although shown as vertical support cables, 21 may comprise any of a number of alternative support configurations. It should further be recognized that similar to the entire suspension system, herein described as supporting compressor motion essentially along a horizontal axis, an evident modification can be provided for vertical or other orientation of the principal axis of motion for the compressor.

The suspension means further comprises a pair of resilient helical discharge tubes 34, axially spiraled about opposite end regions of compressor assembly 20. One end of each discharge tube is rigidly affixed to the outer casing 10 through collar 36, similar to that described for inlet tubing 22. The inboard end of the tubing is rigidly affixed to a discharge port securing collar 42. In other compressor embodiments, only one of these tubes may be used for discharge flow; the other tube can be used for control pressure purposes or any other mass flow requirement. Furthermore, casing 10 need not enclose the compressor assembly, thereby equivalently comprising any suitable support frame.

In the embodiment of FIG. 3, an essentially solid counterbalance mass 38 of known weight is held in a preselected position on discharge tubes 34, 34 by means of a plurality of screws 40. The use of tightening screws or similar clamping means permits adjustment of the position of the mass along the length of the discharge tube. Although described and illustrated as being pre-positioned along the discharge conduit, it should be readily recognized that the counter balance mass may alternatively be positioned along said inlet conduit.

Although not illustrated, the counterbalance mass may alternatively comprise a gas expansion chamber of appropriate mass M.sub.m serving simultaneously as a muffler and counterbalance mass.

Referring now to FIG. 4, from an end view of the compressor system of FIG. 3 it is noted that the helical inlet tubing 22 may spiral from the outer casing 10 to the compressor inlet 28 in a circumferential direction opposite that of the discharge tubing 34. As illustrated in the end view of FIG. 4 the helical inlet tubing 22 spirals in a counterclockwise fashion from the outer casing 10 to the compressor inlet 28. The helical discharge tubing 34 spirals in a clockwise manner from the outer casing 10 to the compressor discharge securing port 42.

Referring now to the vibration isolation characteristics of the suspension system, FIG. 2 illustrates a mass-elastic schematic representation of the suspension system in accordance with the present invention. The following list of terms, when referenced to the cross-sectional view of FIG. 3 will more clearly define the symbols used in the mass-elastic model represented in FIG. 2:

K.sub.1 is equal to the lateral stiffness of that portion of the helical discharge tube(s) 34, 34 extending from a point of rigid attachment to the compressor 42 to a point where the tube engages the counterbalance mass here shown for example as the point 44 where said tube is rigidly clamped within said mass 38.

K.sub.2 is equal to the lateral stiffness of that portion of the helical discharge tube(s) 34, 34, extending from a point of rigid attachment to the outer casing 46,46, or equivalently to a supporting frame, to a point where the tube engages the counterbalance mass at 44,44.

k.sub.3 is equal to sum of two lateral stiffness components. One component is the lateral stiffness of helical inlet tubes 22, 22 extending from a point of engagement to the outer casing shown at 25,25, to a point where the tube (s) engages collar 30 here shown at 31,31. The second component is the lateral stiffness of the vertical support means 21,21.

M.sub.c is equal to the mass of casing 10, or an equivalent supporting frame

M.sub.s is equal to the mass of compressor 20.

M.sub.m is equal to the mass of counterbalance masses 38,38.

K.sub.c is equal to the lateral stiffness of the outer casing's lower support brackets 18;

x.sub.s is the lateral displacement of the mass of compressor assembly 20 from its static equilibrium position;

x.sub.c is the lateral displacement of casing 10 from its static equilibrium position;

x.sub.m is the lateral displacement of the counter balancing mass 38 from its equilibrium position;

w is proportional to the frequency of periodic motion (operating frequency) of the compressor where w=2.pi.f, where f is the operating frequency of the compressor.

All values of stiffness K and displacement x are defined along a horizontal axis, that is, in a direction parallel to the horizontal centerline of compressor 20 and casing 10.

Operation of the dynamic vibration isolator entails the antiphase movement of the counterbalance mass M.sub.m relative to the periodic motion x.sub.s of the compressor mass. Motor and gas-pressure forces generated within the compressor cause the compressor mass to vibrate at operating frequency f. Vibration of the compressor mass causes a spring force to be applied through the K.sub.1 portion of the discharge tube to the counterbalance mass M.sub.m, causing the counterbalance mass to vibrate. This in turn exerts a force through the K.sub.2 portion of the discharge tube acting on casing mass M.sub.c. Concurrently, vibration of the compressor mass M.sub.s applied through stiffness K.sub.3 of the inlet tubes and the vertical support means produces another dynamic force acting on casing mass M.sub.c. By proper selection of the values of K.sub.1, K.sub.2, K.sub.3 and M.sub.m, the forces acting on the casing mass M.sub.c through elements K.sub.2 and K.sub.3 can be made equal in magnitude and opposite in phase such that the resultant or vector sum of these two spring forces acting on the casing will be zero at a particular operating frequency of the compressor. At this condition the motion of the counter balance mass M.sub.m will be antiphase to that of the compressor mass M.sub.s. The actual values of casing mass M.sub.c and casing support stiffness K.sub.c do not influence the values of K.sub.1, K.sub.2, K.sub.3 of M.sub.m as required to achieve zero force transmission at the specified or desired compressor operating frequency.

For a particular size and length of discharge tubing, the position of mass M.sub.m relative to the ends of the helical discharge tube dictates the stiffness values K.sub.1 and K.sub.2. A mathematical preselection of the position of M.sub.m can be calculated. In a working reality, however, such calculations may only closely approximate the correct relationships of stator and counterbalance mass antiphase motion, requiring minor repositioning of M.sub.m. Therefore, in accordance with one aspect of the present invention, the position of M.sub.m (and therein the values of K.sub.1 and K.sub.2), are readily adjusted to provide extremely low transmissibility between the compressor and the outer casing.

For one skilled in the art, a more definitive description of the dynamic vibration isolation means of the present invention is provided by the following analysis of the motions and forces involved. To facilitate illustration, the analysis makes reference to the embodiment including a linear reciprocating motor driven compressor.

The equations of motion for the compressor suspension and fluid flow system are:

Where x denotes linear acceleration (d.sup.2 x)/(dt.sup.2)

Rearranging and substituting for x.sub.s a periodic function of time,

Where w is proportional to the frequency of periodic displacement of the compressor 20, that is, to the operating frequency of the compressor according to the following equation:

Assuming particular solutions of the form

and solving the equations of motion for X.sub.c gives: ##EQU1## X.sub.c will equal to zero when the numerator quantity (-M.sub.m K.sub.3 w.sup.2 +K.sub.1 K.sub.3 +K.sub.2 K.sub.3 +K.sub.1 K.sub.2) is equal to zero and the denominator, which represents the undamped natural frequency equation for the isolation and suspension system, is other than zero. For given values of the various K's and M's, the denominator will be zero for two values of w, which values correspond to the two natural resonance frequencies of the isolation and suspension system; these values will be denoted w.sub.n1 and w.sub.n2. Solving for the condition where the numerator will equal zero, it is found that: ##EQU2## Since, in a physical isolator system, K.sub.3 must be a real, positive number (as must K.sub.1, K.sub.2 and M.sub.m also), Equation 9 dictates that w.sub.n must be less than w. In other words, to obtain zero force transmissibility at a specific compressor operating frequency f, values of K.sub.1, K.sub.2 and M.sub.m must be selected such that

where w and w.sub.n are defined by Equations 5 and 10 respectively.

Once this is recognized, the required value of K.sub.3 can be calculated from Equation 9. The only other stipulation required to have a theoretically valid isolator zero force transmission at operating frequency f is that

where w.sub.n1 and w.sub.n2 correspond to the two undamped natural (resonance) frequencies of the system.

The previous analysis did not account for attendant system damping. From a practical standpoint it is recognized that some damping will be present and that normal manufacturing and assembly tolerances make it improbable that the precise conditions for "zero" dynamic force transmission will be met. Therefore, the following criteria can be used to assure that the actual value of the dynamic force transmitted to the foundation on which the outer casing is mounted will always be small, if not "Zero":

where w.sub.n2 by definition is understood to be the greater of the two system natural frequencies w.sub.n1 and w.sub.n2.

The two system natural frequencies w.sub.n1 and w.sub.n2 can be found from the following equations: ##EQU3##

Employing the previously described design criteria, a refrigeration compressor (similar to that illustrated in FIG. 3 to which reference will be made) was constructed. An electrodynamic linear motor 20B as described in U.S. Pat. No. 3,891,874 is used to drive a linear reciprocating compressor 20 of the type described in U.S. Pat. No. 3,937,600.

Seamless steel inlet tubing 22 having an outer diameter of 5/8 inch and a wall thickness of 40/1000 inch and a total length each of 35 inches are formed into helices having a mean diameter of 9.5 inches and brazed into inlet collars 30 on the compressor housing. The outer end of the inlet tubing is brazed into a conduit coupling collars 26 affixed to the outer casing 10.

Similarly, seamless steel discharge tubing 34 having an outer diameter of 1/2 inches are formed into helices and brazed into compressor discharge ports securing collars 42 and outer casing conduit coupling collars 36. A counter balance 38 of known mass is affixed to discharge tubing 34 at 44. Machine screws compress the axial extremity portions of the counterbalance to rigidly engage the tubing. The elastomeric cushioned lower support member 18 were affixed to a mounting foundation threaded to accept a securing bolt.

The intake and outlet conduit were appropriately connected and the compressor was driven at 230 volts and a line frequency of 60 H.sub.z.

Initial vibration experienced at the outer casing was virtually eliminated after minor adjustment of the position of the counterbalance mass. The elastomeric cushioned lower support (K.sub.c) was removed and the outer casing allowed to rest, unanchored, on the test bench. The motor was restarted and vibration measurements again indicated essentially no vibration of the outer casing and no dynamic force transmitted to the test bench.