US 3588291 A
Description (OCR text may contain errors)
States Patent Inventors Peter W. Cull-wen Ballston Spa; Richard V. Newell, Loudonville, NY. Appl. No. 882,446 Filed Dec. 5, 11969 Patented June 28, 1971 Assignee Mechanical Technology incorporated Latham, NY.
RESONANT PISTON PUMPS 114 C 16 Drawing Figs.
1.1.5. (11 417/417, 417/571 int. Cl ..F04b 35/04, F04b 21/02 Field M Search 417/416, 417, 363, 328, 418, 371
 References Cited UNITED STATES PATENTS 2,721,024 10/1955 Zeh 417/417 3,325,085 6/1967 Gaus 417/416 Primary Examiner-Robert M. Walker Attorney-Joseph V. Claeys ABSTRACT: A resonant piston pump wherein spring means are provided which are secured at one end to the pump housing and at the other end to the pump piston and being operatively arranged and associated with such piston and housing to provide for substantially parallel movement of the piston within but out of contact with the pump cylinder during alternate compression and expansion of the spring means. The piston is driven by an electromagnetic drive means in a resonant mode of vibration with the spring means to provide for maximum power transfer and efficiency of the pump.
PATENTEU JUH28 m1 3,588,291
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SHEET 2 UF 5 [M Mentors We e r Cwuugw Q/W Wowwwy PATENTEDJUHZEHBH 3588.291
SHEET 3 OF 5 I? van toms et-er" Curwen f/C/v'awa The invention described herein relates generally to pumps and more particularly to an improved electromagneticallydriven resonant piston pump having a broad range of applica tions including aerospace, industrial and commercial.
The term pump" is used herein in the broad sense and includes devices or machines that raise, transfer, compress or other wise operate on fluids or attenuate gases. Although, as indicated, the invention has a wide range of applications, it is especially useful in connection with compressors and will be particularly described in that connection.
Many different designs of resonant piston pumps have been proposed for operating with either a liquid or gaseous medium. These designs normally include a cylinder having intake and discharge valves, one or more pistons in the cylinder, and a spring associated with the piston to provide a spring-mass system. To obtain piston movement for pumping a fluid, a solenoid equipped with a movable iron armature is connected with the piston, so that when the solenoid is energized, it produces an electromagnetic force which engages the armature and drives the spring-mass system at the same frequency as the varying electromagnetic field. By designing the resonant frequency of the pump spring mass system to be essentially the same as the frequency of the electromagnetic field, the pump- ,ing work can be performed at high efficiency.
Resonant pumps of the prior art utilizing the above major components generally have not found acceptance in the industrial community, primarily because both the piston and cylinder walls experience excessive wear within a time much shorter than that considered normal for other types of reciprocating pumps. in general, such wear results from the combined effects of a number of independently acting forces which bear on the moving parts. For example, springs or other devices associated with the piston exert forces which fail to keep the piston centered on the cylinder axis during operation. Although the prior art discloses special designs of piston centering devices for use in resonant pumps, none of such disclosures have resolved the problem with any reasonable degree of success. Also, an inability of the solenoid to generate a uniform flux field or to at least cause the armature to always move on the cylinder axis contributes importantly to piston-cylinder wear. Liquid lubricants and piston rings made of different kinds of materials having self-contained lubri cants, also have been employed in minimizing the piston and cylinder wear problem in an attempt to obtain long operating life, but have met with only marginal success. Moreover, such pumps are not suitable for use in environments requiring contaminant-free atmospheres or where the process fluid being pumped must remain free of foreign particles.
To resolve the above and other problems, a major development program aimed at designing and manufacturing an efficient, reliable, gas lubricated resonant pump was undertaken by applicants assignee, with the result that the prior art problems were successfully overcome and outstanding pump performance has been achieved. The completely new design of pump developed is described and claimed US. Pat. Nos. 3,156,405, 3,303,990 and 3,329,334, assigned to the same assignee as the present invention. The major improvements included the development of a new type spring of U shape configuration and in designing the piston outer surface to provide for hydrodynamic or hydrostatic gas lubrication of the piston during its reciprocating movements in the compressor, thus completely eliminating the need for oil or other kinds of lubricants. The U-shaped springs were spaced at 90 intervals with one end connected to the pump housing and the other to the piston so that upon pump operation, the action of the gas and electromagnetic forces acting on the piston, coupled with the spring forces, causes the piston to move on the cylinder axis, thereby eliminating problems of wear caused by rubbing ofthc parts as in prior art constructions.
Since potential applications were envisioned where long life is a necessity, gas bearing technology concepts were applied to the relatively moving parts, thus eliminating the need for solid or liquid lubrication of the moving parts as required by prior art designs. As illustrated in the aforementioned patents, the only moving parts are a solenoid armature, springs, and valve means. Still fewer moving parts are required when the piston inlet and outlet ports are used instead of valve means. Evidence of success of the design resides in the fact that the pressure and flow rates have been maintained at the design point after several thousand hours of testing and trouble-free operation.
Although the new and improved compressors described and claimed in the foregoing referenced patents have been, completely successful, there remains a continuing need for further research and development to provide still more improved high performance, lubricant-free pumps which are not only less expensive and bulky but also exhibit the same or better reliability and efficiency.
Applicants prior resonant pumps have always included design features for lubricating the relatively moving parts with a gaseous medium. Doing so required that substantial consideration be given to the pump design during its early stages to assure having a design of piston, cylinder and springs such that hydrostatic or hydrodynamic lubrication of the piston would result when the pump was placed in operation. Applicants developmental efforts now show that with the correct design and selection of parts, the prior art lubrication problems may be essentially eliminated.
SUMMARY OF THE INVENTION A primary object of the invention, therefore, is to provide an improved resonant piston pump which eliminates the need for lubricants of any kind between the relatively moving parts by maintaining the piston centered in the cylinder during all modes of operation.
Another object of the invention is to provide a resonant spring which when installed in the pump, minimizes the lateral and bending vibrations which heretofore caused the piston to contact the cylinder walls.
Still another important object of the invention is the provision of a highly reliable resonant piston pump of simple, com pact and economical construction.
In carrying out the invention, the need for any lubricant in a resonant pump is obviated by choosing helical springs of the correct material and characteristics suitable for certain sizes of pump, and positioning such springs in a predetermined pattern around the piston. By then connecting the spring ends, respectively to the housing and piston, the lateral and bending vibrations of the spring heretofore encountered are reduced to the point where the piston is caused to move substantially parallel with the cylinder walls thus minimizing contact between the piston and cylinder, and reducing piston-cylinder wear to negligible amounts. By designing the springs to coact with the piston in a particular manner, the size of the pump is substantially reduced while obtaining optimum performance through the use of conventional port or valve means and other pump components. It readily will occur to those skilled in the art that the teachings of this invention are applicable to designs of machines or equipment other than resonant compressors and that the latter is used only to illustrate the preferred embodiments of the invention.
BRIEF DESCRIPTION OF THE DRAWINGS While the specification concludes with claims particularly pointing out and distinctly claiming the subject matter of the invention, it is believed the invention will be better understood from the following description taken in connection with the accompanying drawings wherein:
FIG. 1 is a sectional view in elevation ofonc embodiment of a resonant piston compressor designed in accordance with the teachings of this invention.
FIG. 2 is an end view of the compressor of FIG. 1 with parts broken away to show the arrangement of discharge valves.
FIG. 3 is a perspective view of only the compressor housing illustrating the inlet and outlet passages for the introduction and discharge of gas from the compressor.
FIG. 4 is a plan view of a valve having flexible fingers for controlling the gas discharge passages in the compressor.
FIG. 5 is a sectional view in elevation of a modification of the invention showing a somewhat different arrangement of valves and resonant spring used in the compressor;
FIG. 6 is a plan view ofthe compressor shown in FIG. 5.
FIG. 7 is a view taken on lines 7-7 of FIG. 6 showing the compressor in a different angle of orientation.
FIG. 8 is a view taken on lines 8-8 of FIG. 7 showing the disposition of springs in the compressor.
FIG. 9 is a view taken on lines 9-9 of FIG. 5.
FIG. 10 is a perspective view of the valving arrangement shown in FIGS. 5-9.
FIG. 11 is an enlarged section view taken on lines 11-11 of FIG. 9.
FIG. 12 is a view in elevation, partly in section, of a modification showing a different arrangement of resonant springs used in the spring-mass system of the compressor.
FIG. 13 is a view taken on lines 13-13 of FIG. 12.
FIG. 14 is a view in elevation, partly in section, of another embodiment of the invention.
FIG. 15 is a view taken on the right side of the FIG. 14.
FIG. 16 is a section view of the compressor shown in FIG.
14 taken along line 16-16.
DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS Referring now to the drawings wherein like reference characters designate like or corresponding parts throughout the several views, there is shown in FIGS. 1-4, a resonant compressor in accordance with one embodiment of the invention including a housing 20 having a cylinder 22 and a cylinder head 24 designed for receiving a discharge valve 26. A hollow piston 28 having skirts 30 is adapted for reciprocation within the cylinder, and a machined helical spring 32 designed to a size and configuration to fit within the piston forms part of a spring-mass system as fully described hereafter. The magnetic forces necessary to effect piston movement are generated by a solenoid 34 attached to the other end of the housing, Solenoid 34 comprises a conventional coil 36 wound ona laminated iron core 38 held together by bolts or other fastening means. Anarmature 40 designed for coaction with the solenoid, is connected to the piston 28 by an interconnecting rod or bolt 42. One end of the bolt 42 is attached to the armature at 44 while the other end extends through a hole centrally bored in the piston head and is secured thereto by a nut 46. As shown in FIG. I, the cylinder head cavity 47 is designed to receive and accommodate the projecting nut 46 when the piston reaches the top of its stroke.
Although the material composition of the piston and cylinder usually are different, they nevertheless should have compatible characteristics, such as high dimensional stability, similar coefficients of thermal expansion, good rubbing characteristics and be corrosion resistant. The piston should preferably be made of a material having low mass density so as to minimize stiffness (and thus vibratory force and alternating stress levels) of the resonant spring. In the particular design shown, the piston was made of a carbonaceous material while the cylinder consisted ofa high-nickel content cast iron.
The specific bending and lateral vibration problems encountered in the prior art resulting from utilizing a helicallyshaped spring for achieving resonance, arose out of the use of a spring having a multiplicity of turns with one end attached to or supported by a stationary member and the other end connected to or otherwise associated with the piston. Depending on the particular compressor design, as the spring was compressed and extended during the compression and expansion strokes, the resulting spring forces produced a moment on the piston which caused the piston to simultaneously rotate about a transverse axis and to laterally deflect. This caused contact and consequent wear of the piston and cylinder wall surfaces during operation.
To overcome this problem in the present invention, a helical spring is provided having upper and lower end plates with three or more separate springs interposed therebetween with the coils of each spring lying on the cylinder axis and in the same cylindrical plane, thereby achieving symmetry in the structure. The complete spring assembly was machined from solid stock such that the opposite ends of each spring were integrally formed with upper and lower plates respectively and the ends were spaced 120 from each other on each plate. Since these three coils are symmetrically spaced and located between the end plates, upon application of a compressive or expansive force the spring forces will be axis symmetrically distributed on the end plates, and rotation or lateral deflection of the piston in the cylinder therefore will not take place.
Referring again to FIG. 1, the spring 32 includes three separate and parallel spring coils 50, 51 and 52 interposed between the end plates 48 and 49 to form the integral spring assembly. The ends of the three separate coils merge into their respective end plates 48 and 49 at points spaced 120 apart to assure that the resultant of the three spring forces is transmitted through the spring-assembly centerline, thereby assuring that the piston 28 will move parallel to the cylinder walls. To provide both an efficient and'practical means for clamping the spring body in position, the plate 49 is equipped with an integrally formed flange 53 which extends radially outward from the plate 49; the flange being adapted for positioning between the end of the housing 20 and a support member 54 to which the iron core 38 is attached.
By utilizing this kind of arrangement having three or more symmetrically placed spring coils with a corresponding number of ends merged into the spring end plates, the unequal moments resulting from the spring forces are balanced out, thus eliminating the tendency of the spring, and therefore the piston, to rotate or tilt and to move laterally in the cylinder. Although torsional forces tending to rotate the piston a slight amount about the centerline of the cylinder are still present, they do not effect lateral movement or transverse rotation of the piston and contact between the piston and cylinder walls does not occur.
To provide for intake and discharge of gas from the compressor, inlet ports 56 are circumferentially disposed around the cylinder, and an outlet 58, FIG. 3, which communicates with the discharge manifold 70, exhausts the compressed gas to the system. As shown more clearly in FIGS. 1, 2 and 4, the valve consists of a flexible plate 60 having a number of cantilever reeds 62 which are elastically deflected to an open position to uncover discharge openings 64 in the cylinder head when the compression chamber pressure exceeds the pressure in the discharge manifold 70. One end 66 of the discharge valve is anchored in place by a plate 68 secured to the cylinder head 24 by bolts 69, while the other end terminates in projecting reeds 62 which are free for movement in a direction outward from the cylinder head to the open position. Compressed gas flowing from the compression area into the openings 64 and passed the valve, discharges into a manifold 70 and outlet 58 prior to delivery to an attached system.
In unusual applications where long life and maintenancefree operation is a necessity, to further assure that the piston will not rub or otherwise contact the cylinder walls, a multiplicity of gas inlet ports 72 may be incorporated in the cylinder walls for delivery of a gas for hydrostatically lubricating the piston and cylinder walls. Although the gas lubricant may be supplied from a separate source, the embodiment illustrated in FIGS. 1-3 shows an arrangement wherein the lubricant is furnished by the compressor itself. Fluid is taken from manifold 70 through outlet 74 to passages 76 which communicatc with the gas lubricant inlet ports 72. Should the piston be displaced off the cylinder centerline for any reason, a consequent increase in fluid pressure will occur in the areas of least clearance. Likewise, the pressure in the piston-cylinder clearance space on the other side of the piston will decrease. As a result of the differential pressures, the relatively higher pressure fluid on'one side will act on the exposed piston surfaces and keep the piston centered.
The compressor may be made to operate from either a direct current or alternating current source. In the embodiment illustrated herein, a diode is connected with the solenoid coil to furnish a pulsating current for driving the armature and connected piston at a 60 Hz. frequency although such diode is not required in all cases. The required amplitude of the compressor-frequency harmonic driving force can be calculated closely according to conventional design practices. Also, the mechanical components comprising the piston 28, spring 32, armature M) and shaft 42 connecting the armature and piston, constitute a mass-flexure system which is driven electromag netically and designed to resonate at 60 cycles per second. To show relative values, in one compressor design, the total compression power at p.s.i.a. inlet pressure was calculated to be 50 watts. For a 0.42 inch piston stroke and a 60 Hz. compressor frequency, the resulting 60 Hz. harmonic force amplitude was 11.0 lb. It will be evident that the mass-flexure system must be made compatible with the electromagnetic driving force to secure a resonant mode of operation to obtain the generation of maximum power and efficiency in the system.
In operation, upon energizing the solenoid with a suitable AC voltage, the pulsed current flowing in the solenoid coil produces flux which links the armature. An electromagnetic force is thus produced which draws the armature into the iron core of the solenoid. Simultaneous with that action, the spring 32 compresses and reaches the full degree of compression when the armature completes the end of its travel into the core. As the electromagnetic forces commence moving the armature out of the core, the spring starts to expand and the piston moves to cover ports 56 and start compressing the gas trapped in the cylinder. Continued expansion of the spring with consequent movement of the piston compresses the gas in the cylinder until its pressure exceeds that of manifold 70, whereupon the reeds 62 of valve 26 are displaced from their seat and gas from the cylinder is then discharged into the manifold for delivery to the attached system.
During the compression stroke, the piston does not contact the cylinder walls because the separate spring coils 50, 51, and 52 integrally formed with the plates 48 and 49, uniformly exert equal and symmetric forces on the piston causing it to move along the cylinder axis. The resultant force thus exerted on the piston passes directly through its axis and the piston therefore does not tilt in the cylinder. When the gas in compression chamber 22 equals the pressure in manifold 70, the valve 26 closes and the spring 32 which at that time is in an expanded condition, starts moving the piston on the return stroke (to the right as indicated in FIG. 1). At this time, the electromagnetic forces generated by the solenoid coil 36 assists in drawing the armature and the attached piston into its opening in the solenoid core. Continued piston movement uncovers ports 56 to permit the entry of a fresh supply of gas for compression and the cycle is then repeated. During piston movement, gas from the discharge of the compressor, or from a separate source, may be supplied to inlets 72 to furnish a lubrication medium to the piston-cylinder clearance space to help assure maintaining the piston on the cylinder axis in the manner described above.
As indicated previously, a mass-flexure system comprising the spring, piston and armature is chosen to have a natural vibration frequency essentially the same as the frequency of electromagnetic flux generated by the solenoid coil. As these two frequencies coincide during operation, a resonant mode of vibration is established which produces both maximum power and efficiency in the compressor. Successful performance of the compressor manufactured in accordance with the above design has been achieved and such success is primarily attributable to the use ofa multiple coil spring of the design illustrated in the drawings. Even better performance has been achieved by utilizing a gas for lubricating the cylinder walls in the piston-cylinder clearance space since the gas bearing action produced by the relatively moving parts causes the piston to act like a bearing and stay on the cylinder axis while simultaneously producing friction-free operation. Since reliability and efilciency were important factors considered in the design, a combination of inlet ports and discharge valves were used, rather than full porting arrangements, although it is obvious that either one or a combination can be used.
FIGS. 5-11 illustrate another embodiment of compressor in accordance with the invention although the basic components remain the same. The important changes include decreasing the distance between the armature and piston to obtain better stability-and in locating the spring outside the cylinder walls. This minimizes the moments which resultkfrom unbalanced lateral electromagnetic forces which act on the lateral pole faces at the solenoid armature. By decreasing this distance, the compressor can also be packaged into a more compact construction since it permits locating the resonant spring outside the cylinder and reduces the overall length of the machine. Other changes will be apparent as the description proceeds, and referring more specifically to the drawings, the solenoid 34 comprises a coil 36 wound on a laminated iron core 38, and an armature 40. A plate has one end secured to the armature 40 while the other end terminates in an enlarged head 82. A bolt 84 having its head 86 fitted in a depression in the piston surface interconnects the piston 20 with the armature, the arrangement being such that electromagnetic forces acting on the armature move the piston axially to obtain compression of a gas in the cylinder. The resonant spring 32 positioned outside the cylinder is arranged for compression and expansion as the piston moves in the cylinder. To obtain piston movement in a line parallel with the cylinder walls, the foreces tending to compress the spring reaction forces respectively are uniformly applied to the housing and the piston. The uniform application of force is obtained by utilizing three separate spring coils 87, 88 and 89 nested together to form the spring body 32, with their ends attached to an immovable supporting structure 90. The spring ends are circumferentially disposed from each other at l20 intervals. As shown, a
bracket 90 attached to the housing includes a clamp 92 and.
bolt 93 for firmly grasping and holding each spring coil in a preset position. The other ends of the nested spring coils are supported in a movable yoke 94 which comprises a hub 96 held in place between the piston 20 and armature 40 by bolt 84. Three yoke arms project radially outward from the hub and respectively terminate in cylindrically shaped members 95' which serve to anchor the other ends of the spring coils in position. Cylindrically shaped members 95 may also have a clamp and bolt for firmly grasping and holding each spring coil.
It will be apparent that the yoke effectively serves as a connection means between the spring coils, piston and the solenoid armature. Because of the symmetrical arrangement of the spring coils 87, 88 and 89, and the location of brackets 90 on the housing and member 95 on the yoke for holding the spring ends, the forces applied to the spring body during the piston compression stroke, are uniformly transmitted to equally spaced points on the housing, thus causing the piston to move on the cylinder axis. Likewise, the piston is not diverted from movement on the axis during the return stroke by the expanding springs because the spring reaction forces are absorbed uniformly by the spring supporting structure.
The cylinder head 243 closes the upper end of the compression chamber and serves as a base for valves controlling the inlet and discharge of gas to and from the compression chamber. The head is attached to the housing by bolts R00 and is equipped with an inlet I02 and na outlet 104. The inlet preferably is equipped with a filter I06 for removing foreign particles from the gas introduced into the compressor. Both the inlet and outlet openings respectively merge into supply and discharge manifolds 108 and 110. A plate 112 of ringlike configuration is held in position between the cylinder head and the housing by the bolts 100 and its primary function is to serve as a seat for the valves controlling the flow of gas into and from the compression chamber.
As more specifically shown in FIGS. 9, 10 and 11, the plate 112 is equipped with gas supply openings 114 and gas discharge openings 116, the two sets of openings being isolated from one another by a centrally disposed member 118. Although many different designs of valving arrangements are available and may be employed for controlling the supply and discharge openings, the inlet valve shown consists of a fiat leaf 120 having multiple cantilever reeds 121 which elastically deflect to an open position, as shown in FIG. 11, to admit gas when cylinder pressure is lower than inlet manifold pressure. Likewise, the outlet valve 122 is similar and deflects outward to exhaust gas into the discharge manifold when cylinder pressure exceeds discharge manifold pressure. An end of each valve is located in a depression 130 formed in the compressor housing and is immovably fixed therein by the plate 112. The dotted lines in FIG. 11 illustrate the position of both valves when in an open position. Valve stop 126 limits valve fiexure to a predetermined amount.
By utilizing valves of the above design having reeds which elastically deflect, the 'flow of gas through the supply and discharge openings in the plate can be controlled very effec-- tively. The dimensions of each reed are slightly larger than the openings in the plate with which they align to assure complete control over the gas flow to and from the compressor.
In operation, when the solenoid coil 36 is energized, the electromagnetic forces produced moves the armature and the attached piston in a direction to compress gas in the cylinder in the same manner as described in relation to FIGS. l4. The spring coils simultaneously compress and as the piston reaches the outer end of its stroke, the pressure in the compression chamber overcomes the valve 122 biasing forces plus the prevailing pressure in the outlet, and the discharge valve 122 then opens to permit the discharge of gas through the openings 116 prior to flowing through the manifold 110 and the outlet. The spring reaction forces, coupled with the electromagnetic forces acting on the armature then take effect, and cause the piston to travel on its return stroke. In so doing, the pressure in the compression chamber drops and permits the reeds 121 of the inlet valve 118 to deflect to an opening position, thus drawing in a fresh supply ofgas for compression. Throughout the above described operation, the spring coils act to keep the piston moving parallel to the cylinder walls, thus eliminating the need to provide liquid lubrication to the compressor components. The piston-cylinder clearance space is designed to a close tolerance, but it is obvious that gas from the compression chamber will leak by the piston as it reciprocates on the cylinder axis during operation. If necessary, the leakage gas can be used to provide some gas lubrication of the piston by incorporating special profiles, such as tapers, on the cylindrical surface of the piston and/or the cylinder.
As described in connection with the embodiment of FIG. 1, the mass-flexure system comprising the piston, armature and the springs is made to have a natural frequency of vibration which corresponds with the frequency of the electromagnetic forces generated by the voltage applied to the coil to provide a resonant mode of vibration which permits the compressor to operate at its highest efficiency and with the maximum generation of power in the system.
The major benefits or advantages accruing from complete elimination of bearings in the compressor parts include removal of the need for lubrication of any kind in the piston actuation system. Since lubricants are not required, gases can be pumped without concern for contamination which otherwise would occur in compressors lubricated with oil or other lubricating mediums. The complete elimination of all wearing or surface-to-surface contacting of parts not only reduces the number of components needed, but also provides a degree of reliability not attainable in compressors of similar design. The
mechanical efficiency likewise is increased because the only losses are those resulting from material damping in the resonant spring and external windage effects, both of which are very small. Another important advantage is the pressure-flow characteristics of the compressor which can be adjusted almost instantly by simply changing the voltage applied to the drive solenoid to obtain variance in the piston stroke. It is apparent that this capability permits use of the compressor in a wide variety of applications and in any particular installation where it is desired to maintain maximum compression efficiency over a wide range of operating flows and pressures.
The modification illustrated in FIGS. 12 and 13 is substantially different from that previously described, although both designs have many common features. The major difference lies in the selection and location of the springs used in the mass-flexure system.
As indicated previously, it is essential that the spring coils in the compressor be disposed in a manner to obtain movement of the piston on the center line of the cylinder and thereby not run or otherwise contact the cylinder walls. Although the previous modifications included in construction wherein the coils of the spring were nested in one another, the arrangements shown in FIGS. 12 and 13 utilize a multiplicity of springs which are circumferentially disposed inside the piston.
Referring more specifically to FIGS. 12 and 13, it will be seen that the compressor includes a solenoid 34 comprising a coil 36 wound on a laminated magnetic core 38, and an armature which is designed for axial movement in the solenoid when the coil is energized from a power source. The housing 20 for the compressor serves to form a cylinder 22 having a cylinder head 24 including a discharge valve 26 of the type discussed and illustrated in FIGS. 9-11. The housing includes inlet ports 72 disposed around the housing for introducing a gas directly into the piston-cylinder wall clearance space for lubrication purposes in the same manner as that illustrated in FIG. 1.
The mass-flexure system of the compressor includes a connecting rod interconnecting the armature 40 with the piston 28. A nut 152 secures these parts together and the cylinder head 24 is designed with a recess having a configuration complementary to the nut 152' for permitting the piston to utilize the full compression area in the compression space. Obviously, the bolt may terminate in a depression formed in the piston head should that design be more desirable. The connecting rod 150 is equipped with a plate 154 and is arranged to have a multiplicity of springs 156 located on opposite sides of the plate as shown in both FIGS. 12 and 13. One end of each of the springs is fitted around a collar 158 formed on opposite sides of the plate while the other end of the springs fit over similar collars 160, respectively disposed on a base plate 162 and on the inner surface of inner housing 164. The inner housing 164 is immovably positioned within the piston and is equipped with a projecting flange 166 located between the housing 20 and plate 162 and firmly anchored thereto by bolts, not shown.
In operation, it will be seen that as the coil 36 is energized, the electromagnetic forces produced cause the armature 40, connecting rod 150 and piston 28 to reciprocate in cylinder 22. In doing so, gas is introduced through inlet ports 56 and upon upward movement of the piston in the cylinder, gas is compressed and when it reaches a value greater than the pressure in the manifold holding the discharge valve 26 closed, the valve reeds lift and gas is discharged from the cylinder into the manifold for distribution to an attached system. As the piston travels on its return stroke, the valve 26 will close as the manifold pressure exceeds that in the compression space and further movement of the piston eventually uncovers ports 56 for the introduction of a fresh supply of gas for compression.
As in the previous modifications, the mass-flexure system comprising the armature 40, connecting rod 150, piston 28, and springs 156, operates in'resonance with the electromagnetic forces produced by the coil during alternating cycles. When the resonant condition is reached, the compressor operates at maximum efficiency and power in supplying compressed gas to the attached system. In view of uniform distribution of springs 1156 on both sides of plate 154, the piston 28 will move axially on the cylinder center line and not engage or rub the cylinder walls. As in the previous modifications, gas may be introduced into the piston-cylinder wall clearance space to provide a hydrostatic lubricant and help assure that the piston will not contact the cylinder walls.
It will be evident that on the compression stroke, the springs 156 on the upper side of plate 154 will compress while those below the plate will expand from their initial precompressed condition. Although this design of compressor utilizes a multiplicity of springs, the total effect produced is one of providing great reliability in maintaining movement of the piston one the cylinder axis. The construction shown in FIGS. 12 and 113 lends itself to some of the variations in the compressor structure heretofor discussed in this application. For example, the cylinder head may be equipped with both an inlet valve and discharge valve for controlling the admission and discharge of gas from the compression area. I
The modification illustrated in FIGS. 14-16 is similar in operation to that previously described but differs in the changes made to acquire a more compact structure and one that is more economical to manufacture. Such differences include the means for connecting the armature to the piston and spring supporting yoke and in providing a design of yoke and piston to provide a compact structure necessary for decreasing the size of the compressor. As shown, the modification in cludes a housing 22, cylinder 26 and piston 20 adapted for reciprocation therein. A valved inlet 132. and outlet 134 are provided for controlling the supply of gas to the compressor and for discharging it after compression into a closed system. One side of the housing is equipped with outwardly extending projections which serve as a base for supporting arms 136, 138 and M which terminate in a member designed for attachment to the iron core 38 of the solenoid 34.
Referring more specifically to the parts, the piston is of essentially the same design as the modifications previously described and can be made to provide a relatively close clearance between the piston and cylinder wall for piston sealing purposes. The ratio of radial clearance to piston diameter has typically been about 0.0005 in those designs where gas is used for lubricating the piston. The relative motion of the piston with respect to the cylinder wall in such designs is used to generate a self-acting (i.e. hydrodynamic) fluid film bearing action to center the piston in accordance with well known gas bearing principles. Alternatively, instead of utilizing the relatively moving parts for generating the pressure needed for lubrication purposes, a small bleed flow can be taken from the compressor discharge to pressurize the space between the piston and cylinder walls as described in relation to FIGS. llil, thus providing a hydrostatic bearing centering action. In either case, the actual gas which is being pumped is the gas which is used to lubricate the piston. The close clearance between piston and cylinder reduces piston bypass leakage to small values, while the fluid-film caring action eliminates rubbing contact between the piston and cylinder walls. As a result a very long-lived, contamination-free piston sealing means is obtained.
In one particular compressor having a piston radialclearance-to-stroke ratio of 0.005, the fluid-film bearing function was completely eliminated. In this arrangement, the piston was held centered within the cylinder clearance space solely by the spring. Performance of this compressor showed that after more than 4,000 hours of compressor operation, no evidence of performance deterioration or piston wear is detectable. The valves I and 122 controlling the inlet and outlet are the same as that described in relation to FIGS. 9, 10, Ill and include a plate having openings controlled by flexible fingers attached to the body of each valve.
As illustrated, the compressor is contained within a hermetically sealed enclosure 142, the purpose being to provide a system for isolating vibrations during compressor operation and for permitting the use of a convenient arrangement for introducing gas to the compressor and for discharging it into a reservoir before delivery to the external system. To facilitate the entry of gas, a fitting M4 is secured in fluid tight relationship with the enclosure and a flexible tube 146 interconnects the fitting with a racket M7 attached to the housing by bolts 14% which serves as a gas inlet to the compressor. Since the compressor vibrates during the course of performing its compression function, the flexible tube M6 effectively serves as a vibration isolating device, thus minimizing the magnitude of the mechanical vibrations which otherwise would be transmitted to the enclosure M2.
The yoke arrangement consisting of arms I36, I38 and M0 serve the dual function of supporting solenoid 34 and forming a base for anchoring the ends of the spring 32 in a firm position. To provide an optimum supporting arrangement, the three arms are located on one side of the compressor and are held to the housing by means of bolts M9. As seen in FIG. 16, the arms appear only on one side but may be arranged to extend around the compressor if desired. The other ends of the arms terminate in a supporting arrangement which immovably holds the iron core 36 on the axis of the cylinder. As previously described, the spring element 32 consists ofindependent springs 1150, 152, 154 each respectively being secured at one end by a clamp 156 or similar securing means. The upper ends of each of the springs are attached to the housing 22 by a similar arrangement of clamps.
Although resonant compressor designs have been proposed which yield a theoretical balance of the dynamic forces acting on the frame of the compressor, such designs are not generally attractive from a production cost standpoint. In all of the modifications discussed herein, the dynamic frame forces are unbalanced. It is necessary, therefore, to vibration-isolate the compressor from the base structure. To this end, and to obtain effective, yet inexpensive, isolation, the compressor is suspended from the inside of the hermetically sealed enclosure M2. The supporting means comprises a bar 162 welded or otherwise affixed to the inside surface of the enclosure. A pair of springs I64 interconnect the bar with a pair of spaced brackets 166, which as shown in FIG. 16, extend across the compressor with their ends terminating in brackets 168 or similar securing means. The vibratory forces generated in the compressor during operation are transmitted through the brackets to the springs 1164 and bar 1162 where they are ab sorbed and dissipated by the springs and the enclosure 142.
In operation, gas drawn into the inlet is delivered to the compression area through the reed-type valves I20, and after compression, the valves 122 are forced open and gas is delivered to the cavity formed by the enclosure 142. The cavity forms the function of a reservoir and delivers the compressed gas through an outlet 270 located in a side of the enclosure. The compressor operates in the same manner as in the embodiments of FIGS. 5-11. As it vibrates during operation, the springs I68 are caused to expand and contract in response to the vibrations produced by the moving parts. It will be apparent that the compressor can be operated in any attitude simply by providing a multiplicity of springs 163 around the various parts of the compressor.
As disclosed herein, the approach to piston sealing and for preventing the piston from rubbing the cylinder walls, is to provide a very close clearance between the piston and cylinder wall in those designs where a portion of the process gas is used for piston lubrication purposes. The relative motion of the piston with respect to the cylinder wall is used for generating a self-acting, i.e. hydrodynamic, a gas-bearing action to center the piston; or a small bleed flow can be taken from the compressor discharge to pressurize the space between the piston and cylinder walls to provide a hydrostatic gas bearing centering action to center the piston. To insure minimum manufacturing costs, the piston-cylinder clearance is increased, while still maintaining acceptable leakage, and the piston is kept centered in the cylinder by the sole action of the spring connected to the housing and the piston. In either case, a very low-friction, long-life, contamination-free, piston sealing compressor is obtained. In view of the above, it will be apparent that many modifications and variations are possible in light of the above teachings. It therefore is to be understood that within the scope of the appended claims, the invention may be practiced other than as specifically described.
What we claim as new and desire to secure by United States Letters Patent is:
1. A resonant piston pump comprising:
a housing having a cylinder and a piston adapted for reciprocating movement therein;
a fluid inlet and outlet in said cylinder for admitting fluid through the inlet to be operated on by said piston, and after such operation, discharging said fluid through the outlet;
means for controlling the flow of fluid through said inlet and outlet;
coiled spring means anchored at one end to said housing and connected at the other end to said piston, the arrangement being such that said spring means in undergoing compression and expansion as a result of piston movement during operation, causes the piston to move substantially parallel with and out of contact with the walls of said cylinder; and
drive means including a movable member connected with said piston for driving said piston, springs and the movable member in a resonant mode of vibration to thereby achieve the maximum transfer of power and efficiency in said pump.
2. The combination according to claim 1 wherein said piston comprises a head and hollow skirt integrally attached thereto;
a shaft interconnecting said movable member with the piston head for imparting a reciprocal motion to the piston when the drive means is actuated; and
wherein said spring means are positioned within said hollow piston.
3. Combination according to claim 2 wherein said spring means is machined from solid stock material;
said spring means comprising a plurality of separate coils each of which constitutes a spring integrally merged into said piston head atone end and merged into an integrally formed flangelike member at their other ends; and
means securing said flangelike member to said housing for anchoring said spring means firmly in position.
4. The combination according to claim 3 wherein said inlet comprises a multiplicity of ports spaced around the cylinder peripherally; and
wherein said inlet is controlled by a valve positioned in the cylinder head.
5. The combination according to claim 4 wherein said valve consists of a flexible plate having a plurality of flexible reeds which deflect to an open position when the pressure in said compression space exceeds a predetermined value, and closes when the pressure in said compression space decreases below a predetermined valve.
6. The combination according to claim 5 wherein said outlet is connected through passages in said housing for discharging fluid into the piston-cylinder clearance space for lubricating said piston during operation.
7. The combination according to claim 5 wherein said drive means comprises a solenoid and wherein said movable member comprises a armature; and
means interconnecting said shaft with said armature.
8. A resonant piston pump comprising:
a housing having a cylinder and a piston designed for reciprocation therein;
a head closing the upper end of said cylinder to provide an operation area with said piston during operation;
a valved inlet and outlet in said head for controlling the admission and discharge of a fluid to and from said operation area; drive means supported by said housing and including a movable member connected with said piston to cause it to reciprocate in said cylinder;
spring means comprising a plurality of independent coils each having one end immovably attached to said housing and their other ends to an extension on said movable member, said coils being positioned and arranged in a manner to uniformly transmit or absorb forces to or from the piston respectively, to thereby minimize piston lateral thrust forces and cause the piston to move substantially parallel to the cylinder walls;
said piston, movable member and spring means comprising a mass-flexure system designed to a natural vibration frequency corresponding to the driving force frequency of said drive means to drive said pump in a resonant mode of vibration.
9. The combination according to claim 8 wherein said extension on said movable member comprises a yoke having outwardly extending projections; and means securing an end of each of said coils respectively to the corresponding projections on said yoke.
10. The combination according to claim 9 wherein said coils are disposed around and are of a greater diameter then said cylinder.
11. The combination according to claim 9 wherein said drive means comprises iron core having an electrical coil wound thereon; and
wherein said movable member comprises an armature, so
that upon energization of said coil, the electromagnetic forces produced causes said mass-flexure system to move in resonance therewith.
12. The combination according to claim 11 wherein said armature and the iron core opening into which it reciprocates have complementary surfaces of substantially square configuration.
13. The combination according to claim 11 wherein said valved inlet and outlet comprises openings formed in said cylinder head and valves controlling fluid flow therethrough;
each of said valves comprising a flat element equipped with multiple flexible reeds designed to yield in response to the pressure in said compression area for controlling fluid flow through the inlet and outlet; and
a stop in said cylinder head for limiting the amount said reeds can deflect when moved to an open position.
M. The combination according to claim 13 wherein said valves are mounted on a plate secured between said cylinder head and housing.
UNITED STATES PA"EN'I ()FFE'CE CERTIFICATE OF CORRECTION patent 9 Dated June 28, 1971 inventor) Peter W. (Iurwen wt 11 I "W It is certified that error appears in the above-identified patent and that said Letters Patent are hereby corrected as shown below:
line 56 Column 9, line 14, "one" should read on "earing" should read bearing Column 10, line 6, "racket" should read bracket line 51 "270" should read 170 "predetermined valve" should read Column 11, line 58, predetermined value Signed and sealed this 2nd day of May 1972.
ROBERT GOTTSCHALK EDWARD M.FLETCHER,JR.
Commissioner of Patents Attesting Officer FORM F G-W50 (10-69 USCOMM-DC wave-ps9 9 U S GOVERNMENY PRINTING OFFICE: 1969 0-356-331