US 3286913 A
Abstract available in
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Description (OCR text may contain errors)
Nov. 22, 1966 H. w. KAATZ ETAL 3,286,913
ROTARY PUMP Filed July 1.5, 1964 5 Sheets-Sheet 1 INVENTORS HERBERT W. KAATZ LEO TOBACMAN ATTORNE Y5 Nov. 22,1966 H. w. KAATZ ETAL 3,286,913
ROTARY PUMP Filed July 15, 1964 5 Sheets-Sheet 2 q V 2+ /z m INVENTORS HERBERT w. KAATZ BY LEO TOBA MAN ATTORNEYS United States Patent Office 3,286,913 Patented Nov. 22, 1966 3,286,913 ROTARY PUMP Herbert W. Kaatz, Elyria, and Leo Tobacman, Cleveland Heights, Ohio, assignors to Randolph Mfg. Co., Cleveland, Ohio, a corporation of Ohio Filed July 13, 1964, Ser. No. 382,308 18 Claims. (Cl. 230-152) This invention relates to rotary pumps of the sliding vane type, and has utility when embodied in pneumatic pumps, more particularly dry vacuum pumps, wherein its advantages have presently been demonstrated and enjoyed. This invention comprises, among other things, an improvement over and upon the pump design, construction, operation and results of the illustrative type of pump disclosed in our co-pending application Serial No. 185,073, filed April 4, 1962. In our present invention we utilize our prior invention relating to impregnated mechanical carbon pump parts claimed in said application.
It is among the objects of our present invention to increase the fluid displacement, i.e. capacity of a rotary sliding vane pump for a given length and diameter of rotor and for the gross size or envelope of the whole pump. Conversely, it is one of our objects to provide a pump of minimum size for a desired displacement and capacity.
Another object is to improve the life of rotary sliding vane type pumps, including especially the lives of the vanes thereof, in respect to the size and displacement thereof, to reduce the weight thereof and increase the volumetric efiiciency of such pumps. Another object is to provide an improved rotary sliding vane pump with lower frictional losses, diminished internal fluid leakage, and diminished torque and temperature rise in the pumped fiuid for a given increase in pressure between the intake to the outlet of the pump. Another object is to provide a pump of improved displacement and capacity for the size thereof capable of giving an improved pressure differential without deleterious increase in torque and with reduced wear characteristics compared with the best of the presently known and prior pumps. A further object is to provide a rotary pump of the sliding vane type of more simple and more easly constructed design with an improved rotor drive and coupling- Another object is to provide a pump with an improved bore profile to provide maximum volumetric displacement for a given size pump. Other objects are to reduce friction and wear in a sliding vane rotary pump, to provide deeper vane slots in the rotor, to reduce the driving force on the rotor per unit of torque delivered thereto, and to reduce the cost of maln'ng and using the pump.
These and other objects and advantages of our invention will more fully and at large appear from the following description of the preferred form and embodiment of our invention, reference being had to the accompanying drawings in which:
FIG. 1 is a vertical longitudinal section of an embodiment of our invention in a dry vacuum pump.
FIG. 2 is an end elevation of the intake and driving end of the pump, the left end as viewed in FIG. 1, FIG. 1 having been taken on the vertical medium planes of FIGS. 2-5 inclusive.
FIG. 3 is an end elevation of the exhaust end of the pump, the right end as viewed in FIG. 1.
FIG. 4 is a transverse sectional view taken in the staggared planes of the line 44 of FIG. 1, the planes of the view being offset leftwardly a little in the middle of the pump to show the plate which carries the rotor driving pins in full.
FIG. 5 is a transverse section view of said pump taken in the plane of the line 5-5 of FIG. 1 showing the right end of the rotor, vanes, ring and bore in full.
FIGS. 6, 7 and 8 are charts on the same vertically aligned ordnates with respectively different and appropriate abscissae representing respectively in FIG. 6, the value of radial acceleration acting on each vane by, or with the leave of the cam or contour of the bore; in FIG. 7 the radial velocity of the vane tips in their movement toward and from the axis of rotation of the rotor; and in FIG. 8 the radial extension and retraction, i.e. displacement, of the tip of each vane out of and back into its slot for every angular position of the tip of the vane about the axis of the rotor for a complete working stroke of the vane from the center of one sealing zone to the center of the next following sealing zone. In all FIGS. 6, 7 and 8 the ordinates (vertical lines) measure 180 of angular rotation of the rotor through a complete working stroke, i.e. through 180"- of rotation in a two-lobe pump in which our invention is illustrated and embodied.
Referring now more particularly to the embodiment of our invention shwon in FIGS. 1-5 inclusive, the pump P comprises a central, annular, body or ring G having our novel internal bore B, the cam or contour of which controls the motion of the vanes V in the slots S in the rotor RT toward and from the axis of rotation a-a about which the rotor has its rotary motion. The pump body or ring, vanes and rotor are preferably made out of materials described and claimed in our said co-pending application, such as cast iron for the body with the bore being chrome plated, and mechanical carbon for the vanes and rotor parts only one part of which is impregnated with a small percentage, two to 10 percent by weight, of lithium fluoride. We do not exclude unimpregnated carbon parts from our preference Where conditions of little oxygen and water vapor are not to be met.
The ring G has a slightly greater axial length than the rotor RT whereby to space the end plates 10 and 11 from the ends of the rotor, and normal to the axis aa, for free rotation with allowance for thermal rotor expansion and with the minimum practicable clearance to minimize leakage between the smooth, substantially planar, inwardly facing surfaces of the plates and the smooth, normal ends of the rotor. Plate 10 is disposed at the exhaust end of the pump, the right end as viewed in FIG. 1, and has ports corresponding and co-extensive with the exhaust ports 12, FIGS. 4 and 5, which are cast into the ring G, the ports and the plate 10 not being shown because the section of FIG. 5 is taken within the inner surface of the plate. Compressed gases are exhausted from the pump through the ports 12 and the corresponding ports in the plate 10 and into the exahust chamber 14, FIG. 1, which extends all the way around the right end of the pump within the exhaust housing 18, FIGS. 1 and 3, and communicates freely with the external exhaust connection 16.
The plate 11, FIGS. 1 and 4, contains intake openings or ports 13 which overlie the ends of the intake ports 13a of the ring G. As shown in dotted lines in FIG. 4, the intake ports 13a extend circumferentially clockwise beyond the openings 13 and 13a as far as about the points 13b, and the end plate 11 has circular holes through which intake fluid may enter each of the lobes or work ing chambers of the pump after the vanes have passed beyond the points 13b; all of which extends the effective inlet opening to an angular position of about 58 from the center of the seal, i.e. from the horizontal center line as shown in FIG. 4. The left end of the pump comprises the intake and drive housing 20 which embraces among other things the intake chamber 15 and the inlet connec: tion 17 through which fluid is drawn into the pump. The intake and drive housing 20 has four equally spaced tapped holes not shown as such, which however, are aligned with the longitudinally extending holes 9 in the ring G, FIGS. 4 and 5, and corresponding aligned holes 9a in the plate 11 and corresponding holes, not shown, in the plate 10 and housing 18 through which cap screw 19, the heads of which appear in FIG. 3, serve as tiebolts to hold the assembled structure of the parts together.
The large difference between the major and minor bore diameters, FIGS. 4 and 5,'permits the intake and exhaust ports, being located adjacent the minor diameter, to open longitudinally into the axially located intake and exhaust housings and 18, with the advantage that the exterior of the pump takes substantially cylindrical form having its diameter but little greater than the maximum diameter of the bore, FIGS. 1, 4 and 5.
As shown best in FIGS. 1 and 5, the exhaust housing 18 carries integrally and centrally the right circular cylindrical stud 21 the external circular cylindrical surface of which is finished to be truly coaxial of the axis aa of the pump and truly normal to the plates 10 and 11 and the surface of the housing 18 and 20 which bear upon the plates 10 :and 11. As shown in FIG. 1,
the stud 21 extends from the housing 18 almost the a whole length of the rotor RT and has a free, close rotating fit with the circular, cylindrical, internal bore 22 of the rotor. The bore 22 comprises the bearing for rotary motion of the rotor about the axis a-a of the pump, and the stud 21 comprises the journal therefor.
The external surface of the stud is preferably polished and/or chrome plated for minimum practicable friction with the rotor. The housings 18 to 20-are preferably made of cast iron or other stout metal whose surfaces, as
nomenon and provide that a limited amount of air,
which has been in contact with the halide impregnated part, be directed to flow from the exhaust chamber 14 along the journal 21 and within the bearing 22 and through the center of the housing 20 to atmosphere whereby to diminish friction and wear between the journal and the bearing under such conditions. Under less arduous conditions the halide impregnation may be omitted since the long small-diameter journal and hearing 21-22 combination has its own virtue and advantage. As best shown in FIG. 1, the plate 10 has an enlarged central hole 23 which widely clears the stud 21 and also clears the solid structure in the middle portion of the housing 18 whereby to admit exhaust air from the cham ber 14 to the thin annular space between the stud '21 and the bearing 22 for its movement from right to left, as viewed in FIG. 1, through the space. Such movement performs, also, the negative function of preventing foreign fluid or solid matter from entering the bearing 22 .and the space between the journal and the bearing from the intake end of the pump or the prime mover that drives the rotor.
As shown in FIGS. 1 and 5 the hole 23 in the plate 10 is of greater diameter than the roots or bottoms 31 of the vane slots in the rotor as we have depicted by the dotdash circle 24, FIG. 5, of diameter equal to the hole 23. Therefore air from the chamber 14 is free to enter, and is forced into and through, the roots of the slots S and flows right to left, as viewed in FIG. 1, through the pump and to and through the running space between the journal sleeve 29 and bearing bushing 30 described below. This flow is confined to the small space between the journal 29 and bushing 30 and is restricted to less.
than 1% of the. capacity of the pump, and does however tend to maintain air pressure under? the bottoms of the vanes about the same as discharge pressure.
The rotor being journaled for rotation on stud 21 and having its right end, as viewed in FIG. 1, slidably bearing flush upon the plate 10, is driven through its left end by equi-distantly and/or symmetrically spaced metallic pins or pintles 25, preferably three in number, as shown, entering and engaging the rotor in complementary, longitudinal holes 26, all of which are radially spaced from The pins 25 have their leftward, as viewed, ends or headsi secured, as by copper-brazing in a reducing atmosphere, to the, driving disc 28 and stand normal thereto and. parallel to the axis aa whilst the disc is rotated about and normal to said axis by virtue of having a central hearing and driving sleeve 29 secured to its central. aperture and extending again leftwardly, as viewed, normal tothe disc and centered about the axis aa of the pump in and by the carbon bearing bushing 30 in which the sleeve 29 is journalled for free aligned rotation about the axis aa. The bushing 30 is secured and borne in the hous.--
ing 20, FIG. 1, centrally about said axis. As is also shown in FIG. 1, the bushing 30 affords an axial thrust bearing upon the left, as viewed, side of the disc .28 to position the disc proximate but freely spaced from the adjacent end of the rotor.
The sleeve .29 is splined internally to receive, and be driven by, external spines formed on the inner driving coupling element 33 of the flexible driving coupling 35 which also comprises a similar but ou-twardly disposed coupling element 34 whose external splines 36, FIG. 2, are engageable with a prime mover in familiar fashion to which the pump may be attached conventionally by the attaching flange 40. The coupling 35 also comprises that each of the coupling elements 33 and 34 have radial flanges 37 juxtaposed at their adjacent ends; the flanges in turn carrying oppositely facing studs 38 which engage the perforate flanges of the flexible coupling joint 39 made of rubber-like, but heat-resistant material. The elements 33 and 34 are preferably made of mechanical carbon; the element 33 carrying a metallic washer 32 at its leftward, as viewed, end and interposed tightly'between the flange 37 of the coupling 33 and the end of of the sleeve and coupling element. The washer 32 also extends radially outward for low-friction axial thrust contact with the adjacent end of the bushing 30. A small threaded, central, metallic stud 5 is integrated with the coupling element 33 and extends rightwardly and axially therefrom, as viewed in FIG. 1, and extends through a small central opening in the disc 28. A nut 6 on the stud 5 secures the assembly and assembled relationship of coupling element 33 in the splinedengagement with the sleeve 29 and therefore indriving relationwith the disc 28 and the pins 25. As shown in FIG. 1, the journal 21 is shortened to accommodate the end of the stud 5 and the nut 6. The assembly and disassembly of the whole driving and coupling train is quickly and easily done when the housing 20 is removed from the body of the pump and the rightward,
as viewed in FIG. 1,face of the disc 20 along with pins 25, sleeves 27, stud 5 and nut 6 are openly exposed.
As best shown in FIGS. 4 and 5, each of the lobes 50 lobe 51 'lying entirely below the horizontal bi-sec'ting' plane hh of the pump, which plane passes through the axis a--a and bi-sects the diametrically opposite sealing zones or seals 52 on the left, and 53 on the right, viewed in FIG. 4; the same being seen on the opposite sides right and left as viewed in FIG. 5. As will more fully appear in the description of the bore B to follow, each of the seals or sealing zones comprise 20, the middle of which are of circular arc, bi-sected by the plane h-h, and of radius r, which is but very slightly greater than the radius of the rotor whereby to inhibit the movement of air or gas therethrough and with minimum clearance from the rotor consistent with the necessary practicable allowance for relative thermal expansion and contraction of the ring G and rotor RT.
As shown best in FIG. 5, and as suggested above, the roots 31 of the rotor slots S fall radially within the circular hole 23 of the plate 10, and, by virtue of the small diameter journal 21 and rotor bearing 22, the roots of the vane slots come to greater proximity to the axis a-a than has been practicable in prior practice, where the center of the rotor has usually been keyed or splined to a metallic axle to be driven thereby. Our provision of the central journal and bearing for the rotor extending substantially the whole length thereof, taken with the drive through the pintles 25, widely spaced from the axis of rotation, permits not only the deeper vane slots but also the lower speed between the bearing and the journal 21. Our rotor RT made of mechanical carbon with or without the halide, depending upon the condition of use, has a low co-eflicient of friction with the journal 21 and with the small diameter of the bearing offers a new desirable minimum of frictional resistance to rotation and consequently a longer life with negligible wear.
Referring again to FIG. 4, the working cycle of each of the lobes of the pump arediametn'cally equal and opposite as are the fluid pressures whereby to maintain diametric balance. A description of one of the lobes and one of the working cycles therein will suflice for the description of both. Therefore, a working cycle may be read clockwise in FIG. 4, from the horizontal plane h-h in the middle of the left sealing zone 52 around 180 to the middle of the'right sealing zone 53. The angular progress of-any vane tip, from the middle of the left seal around through these 180 will have the designation 0. For example, the angular distance from 6 equals zero to the end of the sealing zone 52 extends for 10", the first 5 of which are of circular arc and the next 5 of which depart slightly to begin vane excursion, wherefor at this point 0 equals 10 and continues for 48 where 0 then equals 58. The compression zone of the lobe therefor begins at 58 and continues to 140 clockwise around the axis aa whereafter the next 30 comprise the exhaust port which closes when theta equals 170. The next 10 of angular motion of the tip of the vane completes the cycle at the middle of the sealing zone 53 where 0 will equal 180;
The bore contour B of our pump may be described generally by the [following equation:
R=e+r wherein R=the radius from the center of the rotor to the bore sunface,
, e=the radial extension of the vane, and
r=the minimum bore radius which is here but slightly greater than the radius of the rotor.
The 'vane extension e of our pump is described and defined sequentially, segment by segment around the bore, as a function of the angular position of the vane by eleven specific equations stated below, all of the general form:
=the particular angular displacement of the rotor and vanes about the axis of rotation of the rotor for each segment of the bore, and k, l, m: and n are constants.
The curve D of FIG. 8 is the sequential plot of the several specific equations for e for each segment as shown in Table 1 below and now to be more fully explained.
The angular position of thevalie isdefined by'the angle 0 as indicated in FIG. 4. The angle 0 is a function of for each segment of the curve described by a separate equation. In our two-lobe pump 0 equals for each lobe. Our approach to the design of the bore seeks to take into account the various forces acting on the vane at each position assumed by the vane throughout each cycle of operation, and seeks to preserve the vane from shock, wear and adversity. By calculating the maximum acceleration of the vane at various positions we are able to construct the acceleration profile C, of FIG. 6. Integration of the equations which describe the acceleration curve C of FIG. 6 as a function of the minimum bore radius r and of the angular position of the vane, gives the vane velocity curve of FIG. 7. velocity curve yields (the end-product of this presentation for present purposes) the vane extension or displacement curve D of FIG. 8 from which the bore of our pump is laid out and constructed. Integration of the curve D, i.e. the third integration of the acceleration curve, between the. point at which a vane closes the inlet port and the next preceding vane, gives the volumetric displacement for a unit of axial length of rotor and vane for each vane for each lobe. Such a curve is not shown, but may be readily constructed as will fully appear.
In FIG. 8 we have plotted as a dotted curve labeled 0, the vane extension curve of our and other prior pumps as exemplified and disclosed in our co-pending application, Ser. No. 185,073. Construction of the pump bore from curve 0 results in the familiar elliptical bore of the prior art. Comparison of curve 0 with curve D shows that our new improved pump as herein disclosed provides surprisingly greater volumetric displacement than our prior pump which, comparatively speaking exemplifies the prior eflort and failure of all comparable prior art pumps known to us. As shown in FIG. 8, the elliptically contoured bore per curve 0 provides a maximum vane extension of only 0.23r over which our new maximum of 031- represents a 30% increase. Any increase in the prior maximum vane extension with an elliptical or similar bore must cause deleterious wear of the vanes due to the necessity of increased radially inward acceleration of the vanes as will more fully appear from our discussion below about this kind of acceleration in our new pump. The bore of our new pump, on the other hand, by promoting maximum vane extension at and to an angular position ahead of the half-way point of the working chamber of the bore, permits our novel increase in maximum vane extension Without a deleterious increase of radial inward vane acceleration and wear or breakage as was implicit in the prior art.
The acceleration profile of curve C of FIG. 6 is constructed in the following manner. Given the conventional formulae: f=ma, and centrifugal force equals. mv /r, then for a vane of given mass, radial acceleration equal v /R and tangential velocity of the center of mass a=3.05 10- R inches/ degree/ degree This relation indicates that our formula for radial acceleration of the vane is independent of the rate of rotor rotation, except, of course, at such low speeds of rotation that friction preponderates wherein we need not be presently'concerned.
Integration of the" 7 In the first half of the sealing area where :0 to see FIGS. 4, 6, 7 and 8, R may be defined in terms of the minimum bore radius r. At this point where h is the radial dimension or height of the vane, and'the center. of mass is at the half-way height of the vane. Empirical data suggests, as we now prefer, that for our pump, h shall equal about 0.7r to take advantage of the great depth of the roots 31 of the vane slots. The depth of the slots S in enhanced by their inclination rearwardly so that the mouth of each slot is ahead of the root 31 in respect to the direction of rotor rotation. In our preferred form herein, the forward or leading face of each slot S inclines rearwardly about 8 from the radius of the rotor at the intersection of the forward face of the slot and the exterior cylindrical surface of the ,rotor, FIG. 4.
Returning now to the formulae: Substitution of the above values for h and R results in the following equation for the theoretical maximum radial acceleration of the vane (absent friction and'other restraining forces) at and in the sealing area where 0:0 to 5, thus:
a: 1.98 10- r inches/deg.
Having derived this basic expression the construction of the acceleration curce C, FIG. 6, may now begin: To
provide a sealing area of 10 degrees, the vane is restrained from acceleration from 0:0 to 0:5 At 0:5 radial outward acceleration and motion of the vane is initiated at the point 60 at the beginning of the curve C. The. acceleration which we prefer increases smoothly and rapidly to a maximum where 0:10 at the point 61. The smoothness of this rate of increase of acceleration is intended and provided to get the vane off smoothly and quickly to a jerk-free. start. a
We prefer that the maximum value of actual acceleration of radially outward motion of the vane due to centrifugal'force acting thereon, as it flies past the intake ports be about 12% less than the theoretical maximum previously derived and thus be stated:
a: 1.75 x 10- r inches/deg.
This reduction supported by present tests and observations, is intended to allow, and does sufficiently allow for the relatively light frictional and other restraining forces acting in the intake area of 0:l0 to 0:50", between the points 61 and 62 on the curve C. Providing a larger acceleration would tend to risk a gap between the vane and bore with'a consequent risk of leakage past the vane andrisk of vane skipping, resulting in high shock loads on the vane and on the rotor with injurious wear and noise. Weprefer to provide this maximum acceleration of the vane uniformly and continuously whilst 0 grows from 10 to 50 to get the maximum safe pump volumetric displacement for a given rotor length and radius by first flinging the vane outwardly as far and as fast as practicable, given the best materials presently known to us and the operating conditions we contemplate for our vacuum pump.
Deceleration of the radially extending vane begins, as we prefer, while the vane is in the zone of the intake sorts and substantially free of pneumatic load. It is part )f our discovery and teaching that greater maximum safe displacement of our pump can be had by bringingthe lelocity of outward extension of the vane down to zero is at the point 63, FIG. 7, before the vane arrives at the ralf-way point in the working chamber of the pump, IS for example, at about where 0:80. Our improved jore also requires that the vane enjoy radialoutward notion and acceleration for as long a time, measured in legrees of rotation, as possible. The 50 position 62for and the one preceding it, between those vanes and the rotor and the bore, is a maximum. In this area pneumatic load and forces on the vane other than centrifugal force now assume significant values at such normal speeds or between 'aboutlOOO r.p.m. to 5000 r.p.m., and assume amoreimportant role. Consequently we have designed the segments of the acceleration curve C between 0:50
and 0:80", between points 62 and 64, withthe purpose of maintaining the load on the tip of the vane, imposed by the bore, at as small a value as possible, consistent with a smooth, rapid, and satisfactory deceleration of the radially outward motion of the, vane in this part of, the stroke. Notice the reduction in vane velocity from maxi mum at about 55 to zero at the point 63 at 0:80. Had
we chosen a value much less than 0:80 for the end of maximum deceleration, point 65, FIG. 6, i.e. zero radial vane velocity, point 63, FIG.'7, i.e. maximum vane extension or displacement, point 66, FIG. 8, then we would be obliged to impose more severe loadson the vane and rotor in the part of the cycle before 0:80 than we.
presently prefernand'fin-d most advantageous. To keep within our presently chosen safe loads, and have 0 less than for the end of vane deceleration, it would require deceleration to begin before 0:50, and this would reduce vane displacement and volumetric displacement of the pump. Should we choose a value for 0 much greater than 80 for the end of maximum deceleration, zero radial vane velocity and maximum vane extension or displacement, as aforesaid, then' it would be necessary to impose larger inward acceleration forces from the here to the vane tip while, the vane is most heavily burdened with the work of pumping, beginning. when the vaneis first exposed to exhaust pressure when 0 is about 82 or 83, see also FIGS. 4 and 5.
From 0:80 to 0:85 we prefer, to ease up on the initial relatively high value of inward acceleration in order to save the tip of the vane from deleterious wear, FIG. 6 showing acceleration changing from 3.5 to ..94 between the points 65 and 67. A significant result of our achievement of maximum extension of the vane prior to the half-way point in the working chamber is that we have a relatively long period for the return trip of the This longer period the pumping part of the cycle where the vane is subjected to the full differential pressure between the inlet and outlet ports. Therefore ourbore imposes lower forces on the vane tip at this time and consequently the vane is pre-' served from deleterious wear while having a greater extension and sweep than has heretofore been believed to be possible.
As seen in FIG..7, at the point 68 where 0:135 the vane has achieved almost its maximum inward velocity against centrifugal force tending to. throw it outwardly.
It thereupon becomes necessary and advisable to begin to reduce the inward velocity :of the vane so it may have zero radial velocity at the seal zone 53. Therefore, at 0: at the point 69, FIG. 6, weprefer to begin decelerating the vane. To do this the acceleration force induced by the bore is changed between the points 69 and 70 to permit centrifugal force to predominate in slowing the vanes inward speed.
At 0:170", point 71, our preferred acceleration profile has caused the vanes radial inward speed to diminish to practically zero, point 72, FIG. 7. In the profile of FIG. 6, the last 10 of vane travel reflects approximately the first 10.
-above calculations and approximations.
The last 10 of vane displacement is very small as seen in FIG. 8. The vane is brought to zero-radial motion in the seal quite gently.
The acceleration profile of FIG. 6 is the result of the Each of the eleven segments of the curve C in FIG. 6 may be described by a simple equation which, upon double integration, results in the eleven equations for radial vane extension 2 of Table 1, and plotted sequentially as curve D of FIG. 8, as follows:
To illustrate the derivation of the above equations, we 7 will treat the segment of the acceleration curve between 6=5 to =10. The curve of FIG. 6 between these limits, i.e. between 60 and 61, may be expressed mathematically as:
wherein .35 if the slope of the acceleration curve of FIG. 6 between the points 60 and 61.
Integration of the above equation with respect to yields an expression for the radial velocity of the vane:
wherein q is a constant. Since at =0, where 0:5", the velocity v of the vane is zero, the constant q is equal to zero.
Integration of the above velocity equation yields an expression for the extension of the vane:
wherein q is a constant. Since at =0 the vane extension e is zero, the constant q is equal to zero.
Thus we have derived the formula indicated in Table I for the radial extension of the vane between 6=5 to 0:10: i
6X104=.35XT%3=.0583T3 The other equations of Table I were derived and their constants evaluated in the same way: the relation between the curves C, T and D having been explained above.
The novel features of our new pu-mp have also novel coaction to bring about the unitary advantages and results of our new pump. The increased vane extension gained by the bore is also gained, or made possible, by the smallness of the journal and bearing of the rotor,
which in turn is facilitated by driving the rotor through the spaced pintles that are removed from the rotor journal and the roots of the vane slots. So also is the disposition of the pintles facilitated by the independent journaling of the driving and coupling'assembly, axially spaced from the intake end of the rotor journal. This arrangement goes along with axially opposite disposition of the intake and exhaust chambers and housings, and the axial flow in the ring G between those chambers and the radial openings of the ports from the lobes into the ring. The coaction of these and the collateral features and structures herein disclosed has its interdependanoe and interplay in the smallness of the substantially cylindrical envelope of the integrated pump in relation to the greatly increased volumetric displacement and capacity thereof per unit of size.
While we have illustrated and described a preferred form and embodiment of our invention, and have taught our best known way to make and enjoy the same, changes, and improvements will occur to those skilled in the art who come to understand and practice our invention without departing from the essential spirit and principles of it. We therefore do not want to be limited in the scope and effect of our patent to the form and manner herein specifically disclosed, nor in any way inconsistent with the progress by which the art has been promoted by our invention.-
1. A rotary sliding vane pump having an axis of rotation comprising a rotor having an axial length and'a right circular cylindrical exterior surface concentric of said axis and a small diameter interior cylindrical surface comprising a concentric hollow bearing extending the whole length of the rotor, non-radial vane slots parallel with said axis and opening through said exterior surface and lying I approximately tangent to said bearing surface and having said rotor coaxially and symmetrically and of approximately the same length as said rotor and having a noncircular bore describing two lobes with said rotor and two diametrically interposed arcuate sealing zones in sliding contact with said rotor and having diametrically 0pposite inlet and outlet ports respectively opening through said bore adjacent said zones and also opening at longitudinally opposite ends of said ring adjacent said zones, said ring having an imperforate exterior of approximately coaxial cylindrical form and with greater wall thickness comprising said ports, inlet and outlet plates slidably engaging the opposite ends of said ring and having ports selectively corresponding to the inlet and outlet ports in the ends of said ring respectively and closing adjacent ports of opposite effect respectively, said intake plate having auxiliary ports opening to the said lobes respectively adjacent said intake ports of said ring, said plates each having an inner central opening smaller than the exterior of said rotor, intake and exhaust housings closing the longitudinal ends of said pump and having air-tight engagement with said plates and each having an annular chamber communicating with the ports in said plates respectively and having external inlet and outlet connections for said chambers respectively disposed within approximately the envelope of the exterior of said ring, said exhaust housing having an integral axial journal for rotatably supporting said rotor for rotation about said axis and disposed in and for substantially the full length of said internal bearing in the rotor, the said central opening in the exhaust plate clearing said journal and opening to said exhaust chamber to the roots of said slots and to said bearing, said inlet housing having an internal hearing aligned on said axis and sealed-oh? from the said inlet chamber in said housing, a driving disc adjacent the central opening of said exhaust plate and having an axially extending hollow journal rotatably supported in said housing bearing and having symmetrically spaced pintles extending into said holes in said rotor and in driving engagement with said rotor, said hollow journal having internal splines, and a driving coupling having splined engagement therewith, the contour of the interior of said bore comprising sequential segments described by a plurality of equations of the following general form:
wherein R=the distance from the center of the rotor to the inner surface of the bore r=the minimum bore radius =angular displacement of the rotor and vanes about the axis of rotation of the rotor for any segment of the contour of the bore, and k, l, m and n are constants, rsaid vanes having height from tip to root approximately equal to the depth of said rotor. slots and equal to about 0.7 r and said vanes having maximum extension into the bore equal to about 0.30 r.
2- A rotary sliding vane pump having an-axis of rotation comprising a rotor having an axial length and a right circular cylindrical exterior surface concentric of said axis and a small diameter interior surface comprising a hollow cylindrical bearing extending the whole length of the rotor, and having vane slots vanes slidable in said slots, and rotor also having axially extending holes symmetrically disposed about said axis and spaced about midway between said cylindrical bearing surfaces and about mid-way between adjacent vane slots, a ring surrounding said rotor coaxially and symmetrically and of approximately the same length as said rotor and having a noncircular bore describing lobes with said rotor and interposed 'arcuate sealing zones in sliding contact with said rotor and having diametrically opposite inlet and outlet ports respectively opening through said bore adjacent said zones, intake and exhaust housings closing the longitudinal .ends of said pump, said exhaust housing having an integral axial journal cior rotatably supporting said rotor for rotation about said axis and disposed in and for sub .stantially the full length of said internal bearing in the rotor, a driving disc having an axially extending hollow journal rotatably supported in said housing bearing and having symmetrically spaced pintles extending into said holes, in said rotor and in driving engagement with said rotor, thecontour of the interior of said bore comprising sequential segments described by a plurality of equations of the following general form: +m
wherein R=the distance from the center of the rotor to the inner surface of the bore r=the minimum bore radius =angular displacement of the rotor and vanes about the axis of rotation of the rotor for any segment of the contour of the bore, and k, l, m and n. are constants. 3. The pump of claim 2 wherein said vane slots are inclined from radii of the rotor with their roots proximate the internal bearing of the rotor and have a depth of about.
0.7 r and said vanes extend out of said slots about 0.3 r to engage the bore where R is a maximum.
4. The pump of claim 2 with means for conducting fluid from said exhaust housing through said rotor to said intake housing bearing.
5. A rotary sliding vane pump having an axis of rotation comprising a rotor having a cylindrical exterior sur- I .zone through which a vane has just moved.
metrically and of approximately the same length, as said rotor and having a non-circular bore describing lobes with said rotor and diametrically interposed arcuate sealing zones in sliding contact with said rotor and having dia-.
metrically opposite inlet and outlet ports respectively opening through said bore adjacent said zones, intake and exhaust housings closing the longitudinal ends of said pump said exhaust housing having an integral axial journal for rotatably supporting said rotor-for rotation about said axis and disposed in and for substantially, the full length of said internal bearing in the rotor, symmetrically spaced pintles extending into said holes in said rotor and in driving engagement with said rotor, and means journaled in said intake housing'for driving said pintles around said axis.
6. The pump of claim 5 with two lobes in said bore wherein the depth of said vane slots is about seven tenths as long as the radius of the rotor and the contour of the bore permits the vanes to extend about three tenths the length of the radius of the rotor at the place of maximum radius of the bore.
7. The pump of claim 6 wherein the place of maximum radius of the bore is appreciably advanced opposite to i r=the minimum bore radius I v 'qa angular displacement ofthe rotor and vanes about the axis of rotation of the rotor for any segment of the contour of the bore, and k, l, m and n are constants.
9. A rotary pump having a rotorwith a radius r and non-radial vane 'slotsand having inwardly and outwardly movable vanes in said slots and having a bore engaging the tops of the vanes, the height of said vanes and the depth of said slots being approximately 0.7 r, said pump having a working chamber within said bore wherein maximum radius of said bore in said chamber is approximately 1.3 r
-and disposed eccentrically in said chamber, and .the inner r=the minimum bore radius '=angular displacement, of the rotor and vanes about the axis of rotation of the rotorand k, l, m and n are constants.
10. A rotary sliding-vane two-lobe pump, comprising a rotor having vane slots, vanes slidably carried in said slots,
.a bore surrounding said rotor and forming diametrically opposed pumping zones with inlet and outlet openings and sealing zones therebetween, said bore contacting said vanes, the minimum radius of said bore occurring at each of said sealing zones, and the maximum radiusof said bore occurring within each of said pumping zones at an angular position less than half-way between said sealing zones in the direction of vane rotation, said bore influencing said vanes up to maximum 'velocity and down to zero velocity between a sealing zone and said less-than-half-way position.
11. The pump according to claim 10, wherein the location of each of said less-than-half-way positions in each lobe is approximately from the center. of the sealing 12. The pump of claim 11 wherein the, vanes gainia maximum radially outward velocity of about 80 times the radius of the rotor times 10* inches per degree of rotation adjacent the place of closing the inlet opening.
13. A rotary sliding vane type pump comprising a body having a two-lobe bore therein with a minimum radius r, a rotor in said bore having a plurality of vane slots, vanes slidably carried in said slots and contacting said bore, pairs of inlet and outlet passageways having opening and closing ends and communicating with said bore, one inlet and one outlet passageway disposed in spaced-apart relationship in each lobe of said bore, the maximum radius of said bore being located at a point proximate the closing end of said inlet passage and remote from said exhaust passage, and said vanes having a maximum radial velocity of extension out of their slots near the closing end of said inlet passage of about 70-80 r 10 inches per degree of rotation of said rotor.
14. The pump of claim 13 in which said vanes have an extension at said point of maximum radius of about 0.3 r.
15. The pump of claim 13 with sealing zones at the ends of each lobe, the said point of maximum radius of said bore being disposed between about 75 and 85 from the center of the sealing zone adjacent the inlet passage of each lobe.
16. A rotary sliding vane pump having an axis of rotation comprising a rotor having an axial length and a right circular cylindrical exterior surface concentric of said axis and a small diameter interior cylindrical surface comprising a hollow bearing extending the whole length of the rotor, said rotor also having axially extending holes symmetrically disposed about said axis and spaced away from said cylindrical bearing surface, a ring surrounding said rotor coaxially and symmetrically and of approximately the same length as said rotor and having an imperforate exterior of approximately coaxial cylindrical form, housings closing the longitudinal ends of said pump, one of said housings having an internal bearing aligned on said axis, a driving disc having an axially extending hollow journal rotatably supported in said housing bearing and having symmetrically spaced pintles extending into said holes in said rotor and in driving engagement with said rotor, the other of said housings having an integral axial journal for rotatably supporting said rotor for rotation about said axis.
17. A rotary pump having radially movable vanes and a bore surrounding and engaging the tips of said vanes the inner contour of a lobe of 180 angular length of said bore having sequential segements substantially described by the following equations:
wherein e is the radial extension of the tip of the vane as plotted in FIG. 8 in terms of r,
r is the minimum radius of the bore,
is the angular displacement of the rotor and vanes about the axis of rotation of the rotor for each of the eleven above segments of the contour of the bore, and
0 is the angular position of the vane tip throughout the contour of the bore from the center of one sealing zone to the other in the direction of vane and rotor rotation.
References Cited by the Examiner UNITED STATES PATENTS 1,804,604 5 1931 Gilbert 103-136 2,165,963 7/1939 Curtis 103-138 2,544,990 3/ 1951 Harrington 103-136 2,590,727 3/1952 Scognamillo 103-144 2,590,730 3/ 1952 Scognamillo 103-144 2,880,677 4/1959 Grupen- 103-135 2,882,831 4/ 1959 Dannevig 103-161 3,191,853 6/1965 Kroeger 230-152 MARK NEWMAN, Primary Examiner. SAMUEL LEVINE, Examiner.
R. M. VARGO, Assistant Examiner.