|Publication number||US3565558 A|
|Publication date||Feb 23, 1971|
|Filing date||Jan 31, 1969|
|Priority date||Jan 31, 1969|
|Publication number||US 3565558 A, US 3565558A, US-A-3565558, US3565558 A, US3565558A|
|Original Assignee||Airborne Mfg Co|
|Export Citation||BiBTeX, EndNote, RefMan|
|Referenced by (15), Classifications (10)|
|External Links: USPTO, USPTO Assignment, Espacenet|
Feb. 23, 11971 TOBACMAN 3,565,558-
ROTARY PUMP WITH SLIDING VANES Filed Jan. 31, 1969 4 Sheets-Sheet 1 f? f 20 asT Z2 INVE/VY'O/P 4E0 TOBACMAA/ 5910 4 SFsPs/w,
Feb. 23, 1971 TOBACMAN 3,565,558
ROTARY PUMP WITH SLIDING VANES Filed Jan. 31, 1969 v 4 Sheets-Sheet 2 11V VENT 0/? ATTOPA/fVJ L. TOBACMAN ROTARY PUMP WITH SLIDING VANES Feb. 23, 1971 4 Sheets-Sheet 5 Filed Jan. 31, 1969 a 47/ 1 saw/x; N940 QQQhNN X/muzra; 72.22 a
Feb. 23, 1971 L. TOBACMAN ROTARY PUMP WITH SLIDING VANES 4 Sheets-Sheet 4.
Filed Jan. 31, 1969 m 0 6 r 4 4 7 W 0 M V v. a
United States Patent US. Cl. 418150 14 Claims ABSTRACT OF THE DISCLOSURE A rotary pump having sliding vanes carried in axially extending slots in a rotor which turns in a two lobe stator. The surface of each lobe has sequential segments contoured to achieve maximum volumetric displacement within a relatively small stator and also to minimize the stresses occurring in the vanes and rotor during vane extension and retraction while moving circumferentially around a lobe surface to reduce wear and to extend pump life. The pump construction permits assembly and securing of the end plates and other components of the pump using only two threaded fasteners turned into diametrically opposite axially extending openings in the stator wall.
CROSS REFERENCE TO RELATED PATENT This invention relates to and incorporates by reference herein the subject matter of US. Pat. No. 3,286,913, granted Nov. 22, 1966.
BACKGROUND OF THE INVENTION This invention relates to an improved two lobe stator contour which minimizes the forces and stresses occurring in the rotor and in the vanes as they travel around the lobes and which also maximizes the volumetric displacement of the pump having a cylindrical stator of given outer diameter.
The various operating parts of sliding vane type rotary pumps are often subject to rather extreme physical stresses, particularly with respect to the vanes and rotor as the vanes reciprocate rapidly in the rotor slots during their excursion around a stator lobe. These stresses can be attributed to several different forces which act on a vane often simultaneously during its reciprocating travel. The forces cause friction, heat, wear, stress, and fatigue, all of which limit the effective operating life of the pump.
The ultimate effect of friction, wear, stress, and fatigue is to cause a pump failure or breakage. An equally important effect is reduction of pump capacity and efficiency as parts and clearances change with wear and temperature. This effect determines the time during which the pump operates within desired limits of efiiciency and high performance. Vane wear proceeds at a gradual rate up until failure and, thus, often reduces the effectiveness of the pump before actual failure occurs.
The stressing of the vanes can be reduced by reducing the size of the stator lobes relative to the diameter of the rotor and, thus, shortening the vane extension. This reduces the volumetric capacity of the pump, however, and is undesirable for applications where weight and size are critical as in the case of pumps associated with airborne flight instruments.
Another problem associated with prior art sliding vane type rotary pumps, particularly pumps used in connection with flight instruments, is that of achieving a relatively high volumetric displacement in a relatively small pump assembly. Usually, such pumps are coupled or otherwise connected to a rotary accessory drive from a reciprocating aircraft engine or the like and design requirements provide only limited space.
3,565,558 Patented Feb. 23, 1971 With due consideration for adequate wall thickness, the outside diameter of the pump determines the maximum diameter of the lobe contour. The pump performance is derived from the nature of the lobe contour that can be fitted within that maximum diameter. In the description below, however, the lobe contour is more conveniently described in terms of its minor diameter.
US. Pat. No. 3,286,913 achieves desired results by providing for maximum vane extension prior to the mid point between vane travel through one lobe. In this way, the portion of the lobe are which causes retraction of the vane is enlarged so that the retraction can take place at a lower radial velocity than the extension. The improved performance of this type of pump derives inter alia from the reduced friction during retraction which is normally when the greatest friction between the lobe surface and the vane tip occurs.
Other factors, however, still continue to limit the effective life of rotary vane pumps of this type and one of these factors which contributes to wear is the occurrence of events causing excessive forces and stresses on the vanes and rotor all approximately simultaneously or at least within a short time interval during the excursion of a vane around a stator lobe. While several of these events occurring separately would not cause excessive stress on the vane and rotor, their simultaneous occurrence affords a substantial magnification of stresses and contributes greatly to the wear and fatigue of the vanes and rotor, thus limiting the effective performance life of the pump. Also, such excessive stress may exceed the maximum allowable stress on the vanes and rotor.
The forces and moments acting on each vane to cause wear and breakage are complex; but, for the purpose of a general analysis, they may be broken down into thirteen discrete forces which fall into four general classifications. These classifications include pressure forces, motion forces, vane tip forces, and rotor slot reaction forces. The various forces of each classification are more clearly defined below:
(1) Pressure forces:
(a) A resultant pressure force acting tangentially to the vane against the portion extending from the vane slot and normal to its sidewalls. This force is attributable to the pressure differential between adjacent pumping pockets, such as when one pocket is sealed and the adjacent pocket is in communication with an inlet or an exhaust port.
. (b) Three parallel radial pressure forces acting parallel to the sidewalls of the vane at the vanes tip and root. These forces are attributable to pressure differences between the space in the vane slot below the vane root and the adjacent pumping pockets on opposite sides of the vane. The two resultant pressure forces at the tip act on opposite sides of the zone of contact between the vane and the surface of the stator lobe.
(2) Motion forces:
(a) Centrifugal force acting at the center of gravity of the vane and proportional to the square of its angular velocity. This force changes with the radial extension and retraction of the vane so that the magnitude thereof will vary according to the plot of the vanes radial extension during its excursion around a stator lobe.
(b) Coriolis force acting tangentially at the center of gravity of the vane opposite to the direction of angular movement during extension and in the same direction of angular movement during retraction. This force varies with the radial velocity of the vane.
(3) Vane tip forces:
(a) A force acting normal to the lobe surface against the vane tip. This force is, in part, a result of the lobe acceleration forces and, thus, varies with the vane position during travel of the vane around a stator lobe. This force may be considered as composed of two components:
(1) The first acts at the tip of the vane and is directed through the center of gravity of the vane.
(2) The other acts at the tip of the vane normal to the first component. This component, referred to herein as the lobe angle force, tends to make the vane lead its slot in the extension part of the lobe and lag its slot in the retraction part of the lobe.
(b) A frictional force acting tangential to the lobe at the zone of contact between the vane tip and the stator lobe and .which is, of course, dependent upon the magnitude of the vane tip normal force 3(a) above.
(4) Vane slot reaction forces:
(a) A pair of opposed reaction forces normal to the side walls of the vane, one of which acts on the vane in the zone of contact between the vane and the outer edge of the vane slot, and the other of which acts on the opposite side of the vane at its root in the zone of contact between the bottom edge of the vane and the opposite side wall of the respective slot.
(b) Frictional forces perpendicular to the reaction forces 4(a) and occurring in response to relative radial movement between the vane sidewalls and the walls of the slot. These frictional forces occur only during periods of radial movement of the vane in its slot and they vary depending upon the magnitude of their respective reaction forces 4(a).
All the forces which do not act through the center of gravity of the vane apply a movement which reverses during the travel of the vane around the stator lobe.
The sense of the moment depends upon the direction and magnitude of the forces defined above. Thus, the vane tends to rock one way or another in its slot when partially or fully extended. Also, movement of a vane into and out of its slot coupled with the related frictional forces leads to erosion of material from the faces of the vanes and also the adjacent faces of the rotor slot until the vanes are free to rock excessively through increasing amplitudes as the pump wears.
For example, in the pump disclosed in U.S. Pat. No. 3,286,913, the moment acting on the vane initially tends to rock the vane forward in its slot as it starts in excursion around a lobe since the lobe angle force is acting generally in a direction to encourage rotation. Also, the frictional force at the vane tip is relatively low due to the outward movement of the vane from its slot and the consequent reduced vane tip reaction force. When a vane of this prior are pump reaches about the 80 point of its excursion around a stator lobe, however, the moment reverses its sense. In other words, the forces change in direction and magnitude to such an extent that the vane suddenly tends to rock in a reverse manner. This is due in part to the reduction of vane radial velocity (i.e., deceleration) and consequent increase in both vane tip normal force and tip friction force. Also, the lobe angle force changes direction to such an extent that it does not tend to tilt the vane forward but rather acts parallel to the sidewalls of the vane.
This event occurs almost simultaneously with the communication of the pumping pocket on the forward side of the vane with the exhaust port. When this pumping pocket first reaches the exhaust port, a sudden pressure pulse occurs which changes the pressure force acting normal to the vane and, thus, increases the moment at about the time it reverses.
The friction forces deveoped between the vane and its slots during its retraction deleteriously magnify the stress in the prior art pump at the time the vane impacts on the trailing side of its slot. As this flip-flop event occurs, the forces between the vane and rotor slot cause substantial aggravating friction forces acting on the sidewalls of the vane adjacent to the respective portions of the slot which apply reaction forces, all this occurring simultaneously with the reversal of the moment acting on the vane.
In summary, the following events occur almost simultaneously at about the point in the prior art pump:
(1) Reversal of direction of the radial movement of the vane relative to its slot;
(2) A pressure pulse due to the opening of the exhaust port;
(3) A reversal of the moment acting on the vane; and
(4) High frictional forces acting on the sidewalls of the vane adjacent the portions of the slot where the reaction forces are applied.
All of these events combine to produce an abrupt shock load between the vane and rotor and, consequently, promote fatigue and wear and also invite fracture of the vanes or of the rotor.
Another limiting factor which has hindered attempts to enlarge pump capacity by enlarging the lobes is that of providing sufiicient cross-sectional area in the stator for longitudinal threaded openings for the fasteners which secure the various parts of the pump such as end plates and the like. In the past, at least four machine screws at each end have been necessary to assure proper sealing engagement between the end plates and the stator all around the perimeter of the stator. The need for sufficient cross-sectional area in the stator to provide passages for fasteners has, thus, inherently limited the cross-sectional area available for expansion of the lobes within a stator of a given diameter.
Still another limiting factor is the amount of available vane material that can be accommodated in a particular lobe contour. This vane material provides the necessary lubricant between the vane tips and the stator lobe surface for long and useful life. Thus, any substantial increase in the vane material that can be provided will extend the useful life of a pump.
The unique pump construction of the present invention, however, satisfies many of the objections and overcomes the limitations indicated above and unexpectedly affords other features and advantages heretofore not obtainable.
SUMMARY OF THE INVENTION It is among the objects of the invention to increase the effective operating life of sliding vane type rotary pumps and to retain a relatively high pump efliciency for a longer period of time.
Another object is to reduce the physical stresses which occur in the sliding vanes and rotor of rotary pumps during the extension and retraction cycle of the vanes as they travel around a lobe of the stator.
Still another object is to reduce the cross-sectional area of the stator required for fasteners, such as machine screws or the like, to secure the various assembled parts of a sliding vane type rotary pump and, thus, to make more cross-sectional area available for pumping fluid.
A further object is to increase the volumetric capacity of a radial vane type rotary pump.
A still further object is to increase the volumetric capacity of a radial vane type rotary pump without increasing either the axial length thereof or the outer diameter of the stator.
Another object is to provide increased amounts of vane material without sacrificing overall size and pump capacity.
These and other objects are accomplished by means of a radial vane type rotary pump having a stator with two lobes having a minor radius r, a cylindrical rotor having axially extending vane slots. The lobes of the stator en gage the tips of the vanes as they reciprocate during rotation of the rotor and the inner contour of each stator lobe has sequential segments essentially described by a plurality of equations of the following general form:
R is the distance from the center of the rotor to the inner surface of the stator.
r is the minimum stator radius.
41 is the angular displacement of the vanes about the axis of rotation of the rotor and vanes for any segment of the contour of the stator; and k, l, m, and n are constants.
The segments include in each lobe, first, segments wherein R progressively increases at a substantially uniformly increasing rate; second, segments wherein R progressively increases at a substantially uniformly decreasing rate; third, a segment of at least ten degrees of are wherein R remains substantially constant; fourth, segments wherein R progressively descreases at a substantially uniformly increasing rate; and, fifth, segments wherein R progressively decreases at a substantially uniformly decreasing rate.
The lobe contour defined above permits, in certain minimum space applications, cross-sectional area in the stator for only two diametrically opposed axially extending fasteners for securing the end flanges and other pump elements in proper axial alignment when they are in assembled relation. Accordingly, in order to assure a tight fit between an end flange and stator, the end flange may be flexed when in assembled relation to provide re silient pressure to assure a tight fit.
BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a longitudinal sectional view through the centerline of a rotary pump embodying the invention, a portion of the rotor being broken away to show the interior of the stator in elevation;
FIG. 2 is an end elevation from the rearward end of the rotary pump of FIG. 1;
FIG. 3 is an end elevation from the front end of the rotary pump of FIG. 1;
FIG. 4 is a fragmentary elevational view of the rotary pump of FIG. 1;
FIG. 5 is a transverse sectional view taken on the line 55 of FIG. 1;
FIG. 6 is a transverse sectional view taken on the line 66 of FIG. 1;
FIGS. 7, 8, and 9 are charts on the same vertically aligned ordinates with their abscissae representing, respectively:
In FIG. 7, the value of radial acceleration for each vane during its travel around a stator lobe;
In FIG. 8, the radial velocity of each vane during its travel around a stator lobe; and
In FIG. 9, the radial extension and retraction; i.e., displacement of each vane out of and back into its slot during its travel around a stator lobe. In all FIGS. 7, 8, and 9, the ordinates (vertical lines) measure 180 of angular rotation in the two lobe stator;
FIG. 10 is a fragmentary transverse sectional view on an enlarged scale taken through the stator and showing a vane at about the 55 point of its excursion around a stator lobe; and
FIG. 11 is a fragmentary transverse sectional view similar to FIG. 10 showing a vane at about the 78 point of its excursion about a stator lobe.
DESCRIPTION OF THE PREFERRED EMBODIMENT Referring more particularly to the embodiment of the vention shown in FIGS. 1 to 6, inclusive, there is shown a rotary vane type pump P having a central annular body or stator S, a rotor R which turns within the stator, a front flange F secured to the inlet end of the stator S, a back flange B which is secured to the outlet end of the stator S, and a drive assembly D mounted in the front flange F for driving the rotor R.
The front flange F is secured to the stator S 'by two machine screws 10 spaced in diametrically opposite relation to one another and adapted to pass through the flange F and into corresponding threaded bores in the stator S. Likewise, the back flange B is secured to the outlet end of the stator S by means of two machine screws 11, also spaced diametrically opposite to one another and which extend through the back flange B into the opposite ends of the threaded bores in the stator S, the screws 11 thus being coaxial with their respective corresponding machine screws 10 at the opposite end of the stator S.
Ordinarily, this manner of fastening the flange B to the stator S would not provide a tight fit since some looseness could be expected at the abutting portions of the stator and flange in the arc between the machine screws 11. With the unique stator lobe contour of the invention, however, there is insuflicient wall thickness or cross-sectional area at any other place in the stator to accommodate additional adequate threaded bores unless the outer diameter of the stator were increased. This, however, Would defeat the purpose of the invention since the increased weight and size would counterbalance the other advantages.
Since the only portions of the stator wall which provide suflicient thickness for the fastener holes are located between the stator lobes at the position shown, a novel technique has been used to assure a proper fit at those abutting portions of the flange B and stator S spaced from the machine screws 11. According to the invention, the machined surfaces of the back flange B which are placed against the machined end face of the stator S are first hard chrome-plated and then treated with a zinc manganese oxide conversion coating. Following this treatment, the abutting portions of the back flange B adjacent the holes for the machine screws 11 are ground down slightly to remove the conversion coating so that the other abutting portions spaced from the fasteners 11 are about eight ten-thousands (.008) of an inch higher. Thus, when the flange is tightened down to the respective end of the stator S, it first engages the stator at those portions spaced from machine screws 11 and the continued tightening causes a flexing of the flange B to maintain a tight fit. The desired amount of flexing, of course, depends on the dimensions of the parts and physical properties of the material.
The back flange B has a central stud which extends into and substantially through the stator S to provide a journal 12 for the rotor R. The forward end of the rotor R rests against an inlet plate 13 of annular form interposed between the front flange F and the stator S while the opposite end rests against a floating discharge end plate 14, also of annular form and which will be more particularly described below.
The rotor R has a central bore which receives the journal 12, and which provides a bearing for rotary motion of the rotor about its axis. The external surface of the journal 12 is preferably polished or chrome-plated for minimum practical friction with the rotor. The rotor R is also provided with six circumferentially spaced vane slots 15 canted slightly from a radial direction and extending longitudinally the full length of the rotor. Each slot 15 receives a vane 16 which slides in and out during a pumping cycle in each stator lobe. Each vane 16 is preferably provided with a metal jacket 17 such as is disclosed in US. Pat. No. 3,398,884. The jacket is not essential to the invention, however, and the "benefits and advantages still accrue without it.
Referring to FIGS. 5 and 6, it will be seen that the stator S is provided with two symmetrically opposite lobes 18 and 19, the surfaces of which act as cams that regulate the two extension and retraction cycles for the vanes 16 during each rotation of the rotor R. The longitudinal spaces defined by adjacent vanes 16, the rotor R, the surface of a stator lobe, and the end plates 13 and 14 serve as pumping pockets which are moved from an intake zone to an exhaust zone to accomplish the pumping action. Fluids enter the pump through an inlet fitting in the front flange F and pass to an annular inlet chamber 21, also within the front flange F correspondingly, the fluids are exhausted through an outlet fitting 22 in the back flange B which communicates with an outlet chamber 23, also formed in the back flange B.
The entering fluids pass from the inlet chamber 21 to one of two longitudinally extending inlet passages 24 in the stator S which extend from end to end therethrough. Each inlet passage 24 communicates with the pumping pockets in the stator lobes through a series of spaced slots 25 formed in the wall of the stator (FIG. 1). The inlet end plate 13 has two inlet ports 26 which permit passage of the entering fluid from the inlet chamber 21 to each of the two inlet passages 24 and, thence, to the pumping pockets.
The fluid is exhausted from the pumping pockets through another series of spaced slots 27 in the stator wall which communicate with two longitudinally extending exhaust passages 28 on opposite sides of the stator S. The floating discharge end plate 14 is provided with two outlet ports 29 to permit passage of compressed gases from the two outlet passages 28 in the stator S to the outlet chamber 23 in the back flange B.
The discharge end plate 14 is arranged to float in the back flange B in an axial direction. A helical spring 31 bears between the interior surface of the discharge end plate 14 and the back flange B and urges the end plate 14 against the end of the stator S to provide an end seal for the pumping pockets. The floating characteristic of the end plate, however, is not essential to the invention.
As shown in FIG. 1, the marginal interior face of the back flange B is spaced from the respective end of the stator S to accommodate some axial movement of the exhaust end plate 14. The space between the back flange B and the stator is sealed by a positioning ring 33 which tightly clamps against adjacent circumferentail corner grooves 34 and 35 formed in the stator S and back flange B, respectively, as best shown in FIG. 1. The interior face of the back flange B is spaced from the end of the stator S by means of positioning lugs 36 spaced 90 apart around the circumference of the back flange and which bear against the end of the stator S. The lugs 36 are received in notches 37 in the end plate 14 and, thus, serve to position the plate 14 with its respective outlet ports 29 aligned with the respective outlet passages 28 in the stator S.
It Will be noted that the inlet end plate 13 on the other hand is tightly positioned between the forward end of the stator S and the front flange F. A positioning ring 38 similar to the positioning ring 33 serves to seal the circumferential joint. The machine screws 10 pass through openings in the front end plate 13 as shown in FIG. 6 and, thus, inhibit any rotary movement of the plate relative to the stator S or front flange F.
The drive assembly D for coupling the rotor to a rotary drive is best illustrated in FIGS. 1 and 6 and includes a vibration dampening coupling having a driving member 41 with an outwardly extending splined shaft 42 adapted to mesh with a splined socket (not shown) from a rotary drive. The member 41 has a radial flange 43 with rearwardly extending studs 44 that engage a vibration dampening coupling element 45.
The element 45 is best described in US. Pat. No. 3,379,135; and, since it forms no part of the present invention, the operation thereof will not be discussed in detail herein. Briefly described, the coupling element 45 has an axial bore with a contoured pin 46 located therein and also has radial flanges 47 and 48 on both ends for connection to the driving member 41 and an essentially identical driven member 49. The driven member 49 has a radial flange 50 with studs 51 and also a splined shaft 52 received in a splined socket 53 of a tubular shaft 54. The shaft 54 is journalled in a sleeve bearing 55 mounted in the front flange F and has a disc 56 brazed to its inner end as best shown in FIG. 1. A machine screw extends through a central bore in the disc 56 and into a threaded opening in driven member 49 to prevent axial movement of the driven element relative to the tubular shaft 54. Mounted on the disc 56 are six uniformly and circumferentially spaced drive pintles 57 which extend into corresponding circumferentially spaced openings 58 in the rotor R. Each of the drive pintles 57 has a rubber sleeve 59 located thereon to afford a cushioned drive relation between the pintles 57 and the rotor R.
It will be understood that the pump P shown herein is adapted for assembly to accommodate either clockwise or counterclockwise rotation. If the unit is to be assembled for rotor rotation in the reverse direction, the rotor R and stator S are reversed endwise and each of the respective end plates 13 and 14 are reversed to obtain proper registry with the respective inlet and exhaust passages 24 and 28 through the stator S. This is easily accomplished since the two fastener holes in the inlet end plate 13 are spaced exactly 180 from one another and, thus, are diametrically opposed. Also, the notches 37 in the member 14 for the positioning lugs 36 in the back flange B are spaced exactly from one another and can thus accommodate reversal of the end plate 14.
STATOR LOBE CONTOUR For the purpose of illustration, the stator lobes 18 and 19, their relationship to the sliding vanes 16 during the excursion of each vane across a lobe, and the resulting pumping of air from the intake end to the exhaust end of the pump will be described with reference to a clockwise rotation as viewed in FIG. 5. Referring to the pump lobes 18 and 19, each lobe extends around the cylindrical stator between diametrically opposed sealing zones 61 and 62 at opposite sides of the stator. It will be seen that the threaded holes for the machine screws 10 and 11, respectively, are formed in the stator wall adjacent the two sealing zones 61 and 62. Since the lobes 18 and 19 are diametrically equal and opposite, they maintain a balance of forces during the pumping cycle and, thus, avoid any unbalanced forces acting on the rotor journal and bearing.
The angular progress of any vane tip from the middle of the left sealing zones 61 as viewed in FIG 5, around the 180 extent of the lobe will be designated as 0. For the purpose of illustration, the contour of each lobe will be described with reference to (1) a radius from the center of the rotor to the bore surface identified by the designation R, (2) the radial extension of the vane from its slot which will be identified by the designation 2, and (3) the minimum bore radius which is essentially the same as the radius of the rotor and which is identified by the designation r. Using these designations, the vane extension 2 is described and defined sequentially, segment by segment, around the bore as a function of the angular position of the vane expressed by the following formula:
Wherein 5 is the particular angular displacement of a vane 16 about the axis of rotation of the rotor and vanes for each segment of the stator surface beginning with 0 at the counterclockwise limit of the segment, and k, I, m, and n are constants.
The acceleration profile of FIG. 7 is determined by calculating the maximum acceleration of the vane at various positions. Integration of the equations which describe the acceleration curve as a function of the minimum bore radius R and of the angular position of the vane gives the vane velocity curve of FIG. 8. Integration of the velocity curve yields the vane extension curve of FIG. 9 from which the lobe contour of the pump is constructed. Integration of the curve of FIG. 9, or in other words a 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 the rotor and vane for each vane of each lobe. Such a curve is not shown, but may be readily constructed.
The derivation of the specific formulae for each segment of the lobe is accomplished using the desired acceleration curve in a manner well known in the art and as generally described in US. Pat. No. 3,286,913. The acceleration profile of FIG. 7 is a result of these calculations and approximations. Each of the thirteen segments of the curve of FIG. 7 may be described by a simple equation which, upon double integration, results in the thirteen equations for radial vane extension e as set forth in Table I and as plotted sequentially as the curve of FIG. 9.
6 0, 1 :0, and a=0.
Referring next to FIGS. 7, 8, and 9, the curve of FIG. 9 is the sequential plot of the general equation with specific values for each segment of a lobe. The angular position of the vane between the limits of the segment is defined by the angle being equal to zero at the counterclockwise limit of each segment. According to the stator lobe contour of the invention, the resulting velocity and acceleration of the vane during each cycle is such that the occurrence of the most severe stresses acting on the vane are separated from each other with respect to time. This separation of peak stresses and peak moments is accomplished by, first of all, permitting a relatively high velocity of the vane as it extends from its slot during its travel through the first portion of the lobe. Friction between the vane and the surface of the lobe during this portion of the cycle is relatively small since the contour of the lobe is such that it accommodates rather than opposes the centrifugal force causing the vane to extend. This situation changes, however, as the extension velocity decreases prior to reaching the point of maximum extension.
During the latter portion of the extension movement of the vane, i.e., when the velocity of extension decreases from a maximum level to zero and the rate of decrease or, in other words, the deceleration is rather rapid as indicated in FIG. 7. Also, the inlet port closes to the pumping pocket ahead of the vane during the portion of the cycle when equals about 58.
The velocity of extension is reduced to zero at about the 73 point of vane travel around a lobe as indicated in FIGS. 7, 8, and 9. In order that the retraction of the vane not begin at this point, the lobe contour is formed so that, for an interval of about 21 of rotor rotation, i.e., where the extension remains constant at the maximum point. With the stator lobe contours of the invention, however,
the reversal of the moment acting on the vane does not occur simultaneously with achievement of maximum extension of the vane, but, rather, several degrees later. This event occurs at approximately the 80 point or about 7 later in the vane excursion cycle (see FIG. 1). When the moment reversal does occur, however, the vane is not retracting but, rather, is maintaining a constant extension. Thus, the moment reversal is the only major stress producing event which is occurring at this point. Accordingly, two of the principal stress producing eventsnamely, the achievement of maximum vane extension and the moment reversal-have been separated from one another by about a 6 interval.
The conditions just before and just after the moment reversal are illustrated in FIGS. 10 and 11 which show the rotor R and stator S drawn to a size about twice the scale of FIGS. 1 to 6. In FIG. 10, one of the vanes 16 is shown at about the point of its excursion around a stator lobe 18; while, in FIG. 11, the same vane is shown advanced to about the 78 position. In other words, with respect to the vane shown in FIGS. 10 and 11, 0 is equal to 55 and 78, respectively.
Referring to FIG. 10, it will be seen that the normal force indicated by the arrow marked 3a acts in a manner tending to rock the vane forward relative to its slot. Due to the magnitude of this force, the moment acting upon the vane is in a clockwise direction as viewed in FIG. 10. As the vane reaches about the 78 point of its excursion around the stator lobe, the direction of the normal force changes until it acts almost parallel to the sidewalls of the vane. Thus, the effect of the normal force on the moment acting on the vane is considerably reduced over that of FIG. 10. Referring to FIG. 11, it will be seen that an additional force influencing the moment acting on the vane is the tip friction force 31). The tip friction force acts in a manner tending to cause a counterclockwise moment acting on the vane and it has been determined by experiment that the reversal occurs at about the 78 point of vane excursion. Referring again to FIG. 11, it will be seen that this moment reversal occurs after the vane has moved about 6 further around the stator lobe from the point Where its maximum radial extension first occurred, maximum radial extension occurring at about the 73 point.
The next event which occurs during the interval of constant vane excursion is the opening of the exhaust port in the stator wall with respect to the pumping pocket on the leading side of the vane. This occurs as the leading vane reaches about the 147 point. Since the vanes are spaced apart relative to the rotor, the vane at the trailing end of the pumping pocket will experience the pressure pulse described above at about the 87 point which is about 7 after moment reversal.
Vane retractions begin at about the 94 point. The initial retraction movement coincides with the occurrence of the highest frictional forces between the vane sidewalls and the rotor; however, since other stress producing events have already been completed at this point, the effect is minimized.
The retraction movement occurs with a minimum negative acceleration for the vane in order to minimize the force acting at the vane tip tending to push the vane back into its slot. This acceleration is maintained at a generally uniform level until about the l39 point of the stator lobe. At this point, the rate of retraction velocity begins to decrease until the vane is fully retracted at about the 177 point.
While the invention has been illustrated and described with respect to a specific embodiment thereof, it will be understood that variations and modifications may be made in the form and arrangement of the parts and embodiments thereof without departing from the spirit of the invention. The invention, therefore, is not to be limited to the patricular structure and mechanisms herein shown l l and described nor in any manner inconsistent with the extent to which the progress in the art has been advanced by the invention.
1. A rotary fluid pump having a rotor with sliding vanes and mounted for rotation about an axis, and a stator with two lobes, the maximum radius of each stator lobe occurring in a lobe segment of constant radius with respect to the axis of rotation of said rotor and extending at least about 10 of vane rotation around the respective lobe, the midpoint of said segment being at an angular position less than half way around the lobe in the direction of vane rotation.
2. A rotary pump having a rotor with sliding vanes and a stator with two lobes engaging the tips of said vanes, the contour of each of said stator lobes having sequential segments substantially described by a plurality of equations of the following general form:
R is the distance from the center of the rotor to the surface of the stator lobe,
r is the minimum stator lobe radius, and
g is the angular displacement of the vanes about the axis of rotation of the rotor and the vanes for any segment of the contour of such stator lobes, and
k, l, m, and n are constants,
said segments of each lobe including, a first segment wherein R progressively increases at a substantially uniformly increasing rate, a second segment wherein R progressively increases at a substantially uniformly decreasing rate, a third segment of at least about of arc wherein R remains substantially constant, a fourth segment wherein R progressively decreases at a substantially uniformly increasing rate and a fifth segment wherein R progressively decreases at a substantially uniformly decreasing rate.
3. A rotary fluid pump as defined in claim 2 wherein R is at its maximum length in said third segment.
4. A rotary fluid pump as defined in claim 2 wherein said third segment is at least of are.
5. A rotary fluid pump as defined in claim 2 wherein the midpoint of said third segment is less than halfway around its respective lobe in the direction of vane rotation.
6. A rotary fluid pump as defined in claim 5 wherein the midpoint of said third segment occurs between about the 82 and 85 point of vane rotation around the respective lobe.
7. A rotary fluid pump as defined in claim 2 wherein said third segment extends from about the 73 point to about the 94 point of vane rotation around the respective lobe.
8. A rotary fluid pump as defined in claim 2 wherein said successive ones of said segments are separated by transitional segments, said transitional segments comprising a total of less than about 50 around the respective lobe.
9. A rotary fluid pump as defined in claim 8 wherein each of said transitional segments is no less than about 2 around the respective lobe.
10. A rotary fluid pump as defined in claim 2 wherein said stator is of generally cylindrical form and has a front flange and a back flange secured thereto at its opposite ends, each of said flanges being secured to its respective end of said stator by two axial fasteners, each of said fasteners being secured into said stator within said generally cylindrical form between the two lobes.
11. A rotary fluid pump as defined in claim 10 wherein said stator has an essentially planar annular end face against which said back flange abuts at locations adjacent said fasteners and at other locations circumferentially spaced from said fasteners, the abutting portions of said 12 back flange at said locations adjacent said fasteners bein spaced slightly from the plane of the abutting portions at said other location before said fasteners are tightened, whereby when said fasteners are tightened said back flange is flexed inwardly toward said stator at the location adjacent the fasteners.
12. A rotary fluid pump as defined in claim 10 wherein said stator and at least one of said flanges have abutting faces at locations adjacent said fasteners and at other locations circumferentially spaced from said fasteners, one of said abutting faces at said locations adjacent said fasteners being spaced slightly from the plane of the abutting faces at said other location before said fasteners are tightened, whereby when said fasteners are tightened said flange is flexed inwardly toward said stator at the location adjacent the fasteners.
13. A rotary fluid pump as defined in claim 2 wherein said stator is of generally cylindrical form and has an end plate and a flange secured together at one end of said stator, said flange being secured to said stator by two axial fasteners, each of said fasteners being secured into said stator within said generally cylindrical form between the two lobes, said flange and said end plate having abutting surfaces at locations adjacent said fasteners and at other locations circumferentially spaced from said fasteners, and said end plate and said stator having abutting faces at locations adjacent said fasteners and at other locations circumferentially spaced from said fasteners, at least one of said abutting faces at said locations adjacent said fasteners being spaced slightly from the plane of said abutting face at said other location before said fasteners are tightened, whereby when said fasteners are tightened said flange is flexed inwardly toward said stator at the location adjacent the fasteners.
14. A rotary fluid pump having a rotor with sliding vanes and a stator with two lobes engaging the tips of said vanes, the contour of each of said stator lobes having sequential segments described by the following equations:
e is the radial extension of the vane tip from the rotor,
r is the minimum radius of the lobe,
is the angular displacement of the vanes about the axis of rotation of the rotor and vanes for each of the thirteen above segments of the contour of the respective lobe,
0 is the angular position of the vane around the respective stator lobe in the direction of rotation.
References Cited UNITED STATES PATENTS 2,738,774 3/1956 Rosaen 103136(R1)UX 2,791,185 5/1957 Bohnhoif et a1.
103136(H1)UX 3,204,566 9/1965 Feroy 103136(R1)UX 14 3,286,913 11/1966 Kaatz et a1. 230152 3,447,477 6/1969 Pettibone 103136 3,481,276 12/1969 Adams et a1. 230--152 5 CARLTON R. CROYLE, Primary Examiner W. J. GOODLIN, Assistant Examiner U.S. Cl. X.R. 418236 Column Patent No.
UNITED STATES PATENT OFFICE Dated February 23, 1971 Inventor (s) Leo Tobacman It is certified that error appears in the above-identified pateni and that said Letters Patent are hereby corrected as shown below:
line 43. change "in" (second occurrence) to --its--.
line 38. change "circumferentail" to --circumferential--.
line 5 of TABLE I, that portion of the equation reading "1690" should read l690. 5-;
line 11 of TABLE I, that portion of the equatic reading 7s39oo 065325 should read equation (4) that portion of the equation reading ".11342" should read --.113426 equation (10) that portion of the equation Signed and sealed this 7th day of September 1971.
EDWARD M.FLETOHER,JR. Attesting Officer ROBERT GOTTSCI'IALK Acting Commissioner of I
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|US7216526||Jun 3, 2004||May 15, 2007||Rapco, Inc.||Method and apparatus for measuring vane wear in a sliding vane rotary pump|
|US20040197206 *||May 16, 2003||Oct 7, 2004||Henderson Timothy H.||Pump with sealed drive area|
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|U.S. Classification||418/150, 418/236|
|International Classification||F01C21/10, F04C2/00, F01C21/00, F04C2/344|
|Cooperative Classification||F01C21/10, F04C2/3446|
|European Classification||F04C2/344C, F01C21/10|