|Publication number||US3272138 A|
|Publication date||Sep 13, 1966|
|Filing date||Feb 17, 1964|
|Priority date||Feb 17, 1964|
|Publication number||US 3272138 A, US 3272138A, US-A-3272138, US3272138 A, US3272138A|
|Inventors||Connoy Eugene N, Hoffmann Richard W, Parkin James W, Whitmore Charles H|
|Original Assignee||Continental Machines|
|Export Citation||BiBTeX, EndNote, RefMan|
|Patent Citations (8), Referenced by (24), Classifications (9)|
|External Links: USPTO, USPTO Assignment, Espacenet|
Sept. 13, 1965 E. N. CONNOY ETAL 3,272,133
VARIABLE VOLUME PUMP WITH PROTECTION AGAINST OVERHEATING Filed Feb. 17, 1964 5 Sheets-Sheet 1 M MW .Iz/ym AZ Jammy 71 1mm W/Yaflmazm Jami WFarZrzzz [arias/5f Wbzlmz'! Sept. 13, 1966 E. N. CONNOY ETAL 3,272,138
VARIABLE VOLUME PUMP WITH PROTECTION AGAINST OVERHEATING Filed Feb. 17, 1964 5 Sheets-Sheet 2 INLET 32 30 OUTLET\9 I9 27 z/ 35 32 fiwm Jayme AZ Jammy Fla/2am W 1Y0 p 1956 E. N. CONNOY ETAL 3,272,138
VARIABLE VOLUME PUMP WITH PROTECTION AGAINST OVERHEATING Filed Feb. 17, 1964 5 Sheets-Sheet 5 United States Patent 3,272,138 VARIABLE VOLUME PUMP WITH PROTECTION AGAINST OVERHEATING Eugene N. Connoy and Richard W. Hoffmann, Minneapolis, Iames W. Parkin, Burnsville, and Charles H. Whitmore, Savage, Minn, assignors to Continental Machines, Hue, Savage, Minn, a corporation of Minnesota Filed Feb. 17, 1964, Serial No. 345,238 1 Claim. (Cl. 103120) This invention relates to pumps, and refers more particularly to variable volume pumps, wherein the volume of fluid delivered by the pump varies automatically in accordance with the requirements of a fluid pressure responsive system supplied by the pump.
While the invention is applicable to variable volume pumps regardless of type, it is especially well adapted to rotary vane-type pumps wherein a rotor having vanes slidably projecting from its periphery rotates about a fixed axis within a ring that encircles the rotor and shifts towards and from concentrici-ty with the rotor as the volume of fluid delivered by the pump varies. At maximum flow, the ring is farthest from concentricity and at no flow the ring is very nearly concentric with the rotor. It never reaches absolute concentricity since a slight amount of fluid always escapes from the pressure zone of the pump despite the fact that fluid does not enter the system through the outlet of the pump. This slight flow is called case drain, and in most vane-type pumps, as it is in the pump shown herein to illustrate one embodiment of this invention, this case drain is used to lubricate the bearings of the pump.
One of the problems encountered in variable volume pumps concerns the dissipation of the heat resulting from the operation of the pump. As long as the system supplied by the pump requires the flow of fluid, there is usually an adequate heat dissipating circulation through the pump, but when the pump is operating at or near no flow conditions and practically all of the driving energy fed into the pump is converted into heat, the danger of overheating and consequent seizing of the pump is very real, particularly in pumps having a high rated capacity.
In pumps which do not have a very high rated capacity, radiation from the external surfaces of the pump body plus the relatively small flow of fluid from the pressure zone as case drain is usually adequate to prevent dangerous overheating. Thus, for instance, rotary vanetype pumps operating at 1800 rpm. and having a nominal output volume of six gallons per minute, exhibited no serious overheating even at extended period of no flow operation.
However, it was found that when the rotor speed of these pumps was increased from 1800 to 3600 r.p.m. with a view toward stepping up the output of the pump from a nominal six gallons per minute to a nominal twelve gallons per minute, the overheating problem became extremely acute. Since the size of the pump, that is, its physical dimensions, was not increased in doubling its speed of operation, there was no more radiating surface available than before. The only hope of preventing overheating thus resided in increasing case drain. To increase case drain in a rotary vane pump more clearance is required between the end faces of the rotor and the port and wear or thrust plates between which the rotor is confined; but when this is done the resulting abnormally high case drain makes the pump very inefficient.
The dilemma which this discovery posed motivates the present invention. As will be recognized, therefore, it is the purpose and object of this invention to not only provide some means of effecting sufficient circulation of fluid from the high pressure zone of a variable volume pump at times when the pump is operating under no flow conditions, but to do this without sacrificing any of the efliciency of the pump.
Stated in another way, it is the object of this invention to provide a variable volume pump wherein heat dissipating flow of fluid from the pressure zone of the pump adequate to preclude overheating is maintained whenever the pump is operating at or near no flow conditions, but wherein the means for effecting such heat dissipating circulation is rendered inoperative whenever the volume of fluid flowing from the pump outlet is sufficient to satisfactorily dissipate the heat that is generated in the pump.
With the above and other objects in view which will appear as the description proceeds, this invention resides in the novel construction, combination and arrangement of parts substantially as hereinafter described and more particularly defined by the appended claim, it being understood that such changes in the precise embodiment of the hereindisclosed invention may be made as come within the scope of the claim.
The accompanying drawings illustrate one complete example of the physical embodiment of the invention, constructed according to the best mode so far devised for the practical application of the principles thereof, and in which:
FIGURE 1 is an end View of a rotary vane-type variable volume pump embodying this invention, parts of said view being broken away and in section;
FIGURE 2 is a sectional view of the pump taken on a plane parallel with the rotor axis and designated by the line 2-2 in FIGURE 1;
FIGURE 3 is a view from the end of the pump opposite that shown in FIGURE 1, and also having parts broken away and in section;
FIGURE 4 is a sectional view through FIGURE 1 on the plane of the line 44;
FIGURE 5 is an exploded perspective view of the rotor and certain adjacent parts of the pump;
FIGURES 6 and 7 are fragmentary detail views of a portion of the rotor, its encircling ring and the confining port and thrust plates, illustrating by comparison between these two views an essential feature of this invention; and
FIGURES 8 and 9 are respectively detail sectional views through FIGURES 6 and 7 on the planes of the lines 88 and 9-9.
Referring now particularly to the accompanying drawings in which like numerals indicate like parts throughout the several views, the numeral 5 designates the body of the pump which comprises a main section 6 and acover 7 bolted thereto. The inlet 8 and the outlet 9 of the pump are in the main section 6 and open downwardly through the mounting base 10 of the pump. The body of the pump defines a hollow chamber 11 in which the rotor 12 of the pump is located, the rotor being fixed to its shaft 13 which is journaled in bearings 14 and 15 located respectively in the main section 6 and the cover 7. The end of the shaft which is journaled in the main section protrudes therefrom to provide for connection of the pump rotor with the shaft of a drive motor, not shown.
The rotor 12 has a plurality of substantially radially disposed slots 16 extending in from its periphery and opening to the opposite ends of faces of the rotor which are fiat. Slidably received in these slots are vanes 17, the outer end portions of which project beyond the periphery of the rotor to span the space between the rotor and the inside surface 18 of a pressure ring 19 which encircles the rotor. Thus the ring 19, the rotor, the vanes and end walls provided by a port disc or plate 20 and a thrust plate or disc 21 between which the assembled rotor, vanes and pressure ring are confined, coact to define fluid pumping chambers 22 which travel in circular orbits as the rotor revolves.
A governor spring 23 which reacts between an adjustable spring seat 24 in the pump body and the adjacent side of the pressure ring 19, yieldingly urges the ring towards its position of maximum eccentricity which, in the pump illustrated, is defined by a fixed abutment or pad 25 on the pump body. The force with which the spring 23 acts upon the pressure ring determines the pressure which the pump will develop and maintain.
Because of the eccentricity between the pressure ring 19 and the rotor, the pumping chambers 22 increase in size during one-half of their orbital rotation about the rotor axis and decrease in size during the other 180 of rotation. As the pumping chambers 22 increase in size or volume, they sweep past and communicate with arcuate inlet ports 26 and 27 located respectively in the port plate 20 and in the thrust or wear plate 21, to receive fluid therefrom; and, as they decrease in size, they sweep past and communicate with an arcuate discharge port 28 which communicates wit-h the pump outlet 9.
In this manner fluid is drawn into the pump through its inlet 8 which, in practice, is connected with the reservoir of the system, not shown, or any other suitable source of fluid, and forced from the pump through its outlet which, of course, is connected with the system to be supplied. It follows, therefore, that both a suction zone to which the inlet leads and a pressure zone at all times communicated with the outlet of the pump, exists inside the pressure ring 19.
As the flow requirements of the system decrease, the fluid pressure inside the pressure ring causes the ring to shift towards concentricity with the rotor in opposition to the governor spring 23, and, in so doing, correspondingly reduces the volume of fluid delivered by the pump. This fluid pressure inside the pressure ring also applies a force on the ring tending to displace it from its desired path of translatory motion towards and from concentricity with the rotor, the direction of this force being upward, as viewed in FIGURE 1, so that the ring is held against an abutment 29 fixed in the pump body. Since the surface on the abutment 29 against which the pressure ring bears is fixed, the ring rolls thereon as it shifts back and forth in consequence of the varying flow demands of the system.
The rolling of the ring on the abutment 29 in conjunction with the other forces acting upon the pressure ring, causes the ring to creep around the rotor and thereby uniformly distribute the wear on the inner surface of the ring which results from the sliding or rubbing engagement of the vanes therewith. This wear distributing feature, however, is no part of the present invention, being instead the subject matter of the copending application Serial No. 337,043, filed January 10, 1964, by Charles H. Whitmore, Russell G. Winquist and Sheldon E. Thorson. As more fully described in that application, the inner ends of the slots 16 in the rotor are successively communicated with the suction and pressure zones of the pump so that fluid pressure acts on the inner ends of the vanes to press them outwardly into firm engagement with the inside surface of the pressure ring. However, and in spite of the fact that the vanes are thus pressed against the inside surface of the pressure ring, some fluid will flow past the vanes from one pumping chamber 22 to the next. This is called slippage, and obviously it has an adverse effect upon the efliciency of the pump.
The flow of fluid from the pressure zone through the running clearance which must be provided between the opposite end faces of the rotor and the port and thrust plates 20 and 21, and which leaves the pump as case drain, also adversely effects the eificiency of the pump, but this flow can be, and in the present pump is, used to lubricate the bearings. Thus, as best seen in FIGURE 4, the fluid which escapes from the pressure zone through the clearance between the port plate or disc 20 and the adjacent side of the rotor, flows through the bearing 14 and then 1. 2138 1 pump through a case drain port 30,
which, like the inlet and outlet, opens downwardly through the base 10 of the pump. By the same token, the fluid which escapes from the pressure zone through the running clearance between the thrust plate or disc 21 and the adjacent side of the rotor, flows through the bearing 15 and, from there, is conducted through passage means 31 to the hollow interior of the pump body to be exhausted therefrom through a passage 32 which leads to the case drain port 30.
The running clearance provided between the opposite faces of the rotor and the port and thrust plates between which the rotor is confined, and hence the amount of fluid which leaves the pressure zone of the pump as case drain, is governed by the thickness of a shim or shims 33 interposed between the thrust or wear plate 21 and the adjacent inner face of the cover 7. Obviously, the thinner the shim or shims, the greater will be the clearance and, consequently, the larger the case drain flow-all other contributing factors being the same. Equally obvious is the fact that if the case drain flow is too large, the efficiency of the group suffers, since fluid which is permitted to leave the pressure zone of the pump as case drain performs no useful work, other than the lubrication of the bearings. Accordingly, the shim thickness should be chosen to provide only the necessary running clearance.
As explained hereinbefore, as long as the system requires fluid flow, the heat which is produced during operation of the pump is adequately dissipated or carried off by the fluid circulating through the pump. Case drain flow, of course, also carries off some of the heat and, in addition, part of it is radiated from the external surfaces of the pump, but as long as the system demands any appreciable flow it is the fluid delivered to the system by the pump which carries off the heat. The problem arises when there is no such flow so that case drain and radiation must be depended upon to prevent overheating.
It is known that a rotary vane type pump operating at 1800 rpm. and having a nominal rated capacity of six gallons per minute, has suflicient case drain and sufficient radiation to preclude overheating, even during extended periods of no-flow operation when practically all of the driving energy imparted to the pump is converted into heat. But when it was discovered that this same pump could be driven at twice the speednamely, 3600 r.p.m., to achieve a nominal volume rating of twelve gallons per minute, provided that cavitation could be avoided, the normal case drain and radiation was unable to dissipate enough heat to prevent serious overheating and very real danger of having the pump seize.
Although it does not constitute any part of the present invention, it should be explained that cavitationwhich was observed when the pump, in exactly the same condition as it had been for 1800 rpm. operation, was simply driven at a higher speedwas overcome by simultaneously bringing the fluid into the pumping chambers from both sides of the rotor. As shown in FIGURE 2, this is accomplished by a crossover passage 34 leading from the inlet 8 through the main section of the body and into the cover 7 which has a cored transfer passage 35, the mouth 36 of which is arcuate and registers with the arcuate inlet passage 27 in the wear plate 21.
The solution of the overheating problem which this invention provides, consists in providing controlled communication between the space inside the pressure ring and an exhaust passage which comprises the hollow interior of the pump body, the passage 32 and the case drain port 30. This is accomplished by the simple expedient of providing a shallow notch or recess 37 in the inner face of the port plate or disc 20 and of the thrust or wear plate or disc 21. In effect, these notches or recesses may be considered ports or months through which the exhaust passage connects with the pressure zone of the pump. The notches or recesses are so located and are of such size and shape that the adjacent flat end faces of the pressure ring extend across them to close these mouths by closing off communication between the notches or recesses and the space inside the pressure ring when the ring is in any position it occupies during any appreciable flow of the pump; but not when the ring is in its position most nearly concentric with the rotor, which position it occupies when no fluid flows from the pump to the system. The proper location for the notches or recesses 37 is thus :on the plane of translatory motion of the pressure ring as it shifts towards and from concentricity with the rotor.
The length of the notches or recesses in the direction of pressure ring adjustment must be so related to the thickness of the ring wall, or-more accurately-the radial width of its flat end faces, that the inner end 38 of the notches or recesses opens to the inside of the ring only when there is no flow or practically no flow from the pump outlet. The outer ends of the notches or recesses are at all times open to the hollow interior of the pump body. This relationship between the size of the notches or recesses and the pressure ring is best illustrated in FIGURES 6 to 9, inclusive.
The notches or recesses 37 and the adjacent flat end faces of the pressure ring thus coact to form valves governing communication between the pressure zone of the pump and the exhaust passage leading from the pump. When these valves are open, the mouths of the exhaust passage are open and fluid flows from the pressure zone of the pump and out of the exhaust passage. The rate at which it flows is greater than normal case drain, as it should be; but obviously, even then, the flow must be restricted or limited. The dimension of the notches or recesses to provide the proper rate of how may be readily calculated. For purposes of illustration, a pump driven at 3600 rpm. and operating at no flow and 1000 psi, must have a circulation of at least 1500 millil-ters per minute (.39 g.p.m.) out of its pressure zone to prevent seizing due to overheating, but only 300 to 400 milliliters per minute when the pump is delivering one or more gallons per minute.
Preferably, the inner ends 38 of the notches or recesses are convexly curved, as shown, rather than square, to efliect a progressively accelerated cessation of heat dissipating flow from the pressure zone, rather than abrupt termination thereof, as the pump begins to deliver fluid to the system.
Although there are two notches or grooves 37 in the embodiment of the invention shown and described, one at each face of the rotor, it is obvious that only one such notch or groove could be used provided it had the required flow capacity.
From the foregoing description, taken in connection with the accompanying drawings, it will be readily apparent to those skilled in this art that this invention provides a simple and practical solution to the overheating problem in variable volume pumps, since it provides for adequate heat dissipating flow or circulation of fluid from the pressure zone of the pump when the pump is operating at or near no flow conditions, while maintaining pressure, but terminates that flow as soon as appreciable flow from the pump to the system supplied thereby begins.
What is claimed as our invention is:
In a variable volume pump of the rotary vane type having (1) a body with an inlet and outlet, and defining a rotor chamber communicating with the inlet and outlet and having opposite side walls,
6 (2) bearing means in the body, ('3) a rotor journalled in said bearing means and confined in the chamber between the side walls thereof, the rotor having vanes slideably projecting from its periphery, and (4) a ring which encircles the rotor and is shiftable towards and from concentricity with the rotor and coacts with the vanes to provide pumping chambers by which fluid is taken from the inlet, pressurized and delivered to the outlet at a substantially constant pressure but in a volume which varies with the position of the ring, the volume being maximum when the ring is farthest from concentricity and zero when the ring is concentric with the rotor,
so that part of the space inside the ring is at all times communicated with the outlet and during operation of the pump constitutes a pressure zone,
wherein some of the fluid which inevitably leaks from said pressure zone flows to the bearing means to lubricate the same, and wherein the body also has a drain port at all times in communication with the bearing means and with the outer portion of the rotor chamber that encircles the ring, to receive and carry off fluid which flows through the bearing means and fluid which leaks out of the pressure zone into the outer portion of the rotor chamber, the improvement which prevents overheating of the pumpdue to conversion into heat of the energy expended to drive the rotor and maintain the pressure when the volume of fluid delivered by the pump is zero or near zero, and which improvement comprises:
means defining a fluid passage of limited capacity to communicate the pressure zone inside the ring with the outer portion of the rotor chamber and hence with the drain port,
said fluid passage having open communication with the outer portion of the rotor chamber and having its mouth in one of the side walls of the rotor chamber in position to be covered and closed by the adjacent surface of the ring whenever the ring is in any position it occupies when the volume of fluid delivered by the pump through its outlet is sufficient to prevent overheating of the pump and to be uncovered and open when the ring is in the position it occupies when the volume of fluid delivered by the pump is at or near zero.
References Cited by the Examiner UNITED STATES PATENTS 2,637,275 5/ 1953 McFarland 103-120 2,649,739 8/ 1953 Huiferd et a1. 103-120 2,651,994 9/ 1953 DeLancey et al 103-120 2,740,256 4/1956 OMalley 103-120 2,768,585 10/1956 Hardy 103-120 2,805,628 9/ 1957 Herndon et a1. 103-120 2,875,699 3/1959 Herndon 10 3-1 20 3,120,814 2/1964 Mueller 103-120 MARK NEWMAN, Primary Examiner.
WILBUR J. GOODLIN, Assistant Examiner.
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|U.S. Classification||418/27, 418/102|
|International Classification||F04C14/00, F04C15/00, F04C14/22|
|Cooperative Classification||F04C15/0096, F04C14/223|
|European Classification||F04C14/22B, F04C15/00G|