EP1421282B1

EP1421282B1 - Fluid displacement pump with backpressure stop

Info

Publication number: EP1421282B1
Application number: EP02728628A
Authority: EP
Inventors: Arthur Vanmoor
Original assignee: Individual
Current assignee: Individual
Priority date: 2001-08-27
Filing date: 2002-03-28
Publication date: 2007-11-07
Anticipated expiration: 2022-03-28
Also published as: ES2295340T3; EP1421282A4; DE60223388T2; WO2003019009A1; DE60223388D1; ATE377709T1; EP1421282A1

Abstract

The fluid displacement pump enables substantially continuous pumping from a low-pressure side to a high-pressure side substantially without any backflow or backpressure pulsations. Liquid or gas is injected to the high-pressure side by way of mutually intertwined worm spindles that form a fluidtight displacement system. The blades (9A, 9B) of the impeller system are slightly curved from the inside out, i.e., from their axles (31) to their periphery, so as to ensure a tight seal between adjacent blades. The orientation of the blades is almost flat, i.e., their attack angle relative to backpressure is close to perpendicular so that they will turn quite freely in the forward direction, but will not be turned backwards by a pressurized backflow. The impeller rotation that is introduced via the spindle shafts (31) nevertheless leads to a volume displacement towards the high-pressure side, for instance, towards a chamber to be pressurized or to be subjected to equal pressure.

Description

The invention relates to a fluid displacement pump, comprising a housing formed with a chamber having a wall defined by two mutually intersecting cylindrical openings defining respective cylinder axes and two axles respectively disposed at and rotatably mounted about respective axes coaxial with said cylinder axes, said axles each carrying a helically rising blade sealing against said wall of said housing and engaging into one another, for pumping liquid and/or gas phase materials.
The fluid pump is useful, as described in my earlier applications, in the context of an output system of an internal combustion engine or a turbine engine and an input system for injecting fluid into the combustion process. The input system, in that case, includes a displacement pump, specifically for use with air and water, which can be utilized as a gas compression pump in the internal combustion engine and the turbine.
Fluid displacement pumps are subject to a variety of applications in engineering. For instance, such pumps are utilized in compression systems such as air compressors and as fluid pumps. For example, British Patent Specification 265,659 to Bemhard discloses an internal combustion engine with fuel pressurization separate from the combustion chamber. There, fuel is pressurized in a compressor and the pressurized fuel is fed from the pump to the engine through a port assembly.
U.S. Patent No. 1,287,268 to Edwards discloses a propulsion system for a motor vehicle. There, a compressor formed with mutually interengaging helical impellers pumps to an internal combustion engine which is also formed with mutually interengaging helical impellers. The internal combustion engine drives a generator, which pumps hydraulic fluid to individual hydraulic motors that are disposed at each of the wheels. The impellers of Edwards are formed with "flat" blades of a constant thickness from the axle radially outward to their outermost tip.
The efficiency of fluid pumps with interengaging impeller blades is dependent on the seal that is in effect formed between the blades. While the outer seal is relatively easily obtained with a corresponding housing wall, the inner seal between the blades, i.e., at the location where the blades overlap is rather difficult to obtain. In the prior art system of Edwards, for example, the flat blades do not sufficiently seal against one another and the corresponding efficiency of the double impeller pump is therefore relatively low. Certain applications of the fluid pump require a better seal and better backflow prevention.
US 1,698,802 describes a fluid displacement pump with a blade on a left hand rotor having a convex curved surface and a blade on a right hand rotor having a concave curved surface. Therefore, engagement of the blades of the two rotors is only possible at the bottom surface in the grooves of the screws and the outside parameter of the screws respectively.
GB 2 182 393 A describes an intermeshing screw pump, wherein the widths of screw threads of a screw at their roots are larger than the widths on the pitch circles. The surface of the blade has a concave rounded surface extending from the peripheral wall of the axle radial outwardly and merging into a convex rounded surface at the outer periphery of the blade.
DE 297 20 541 U1 shows in principal a blade with the same shape as described in GB 2 182 393 A while in the DE 199 41 787 A1 the surface of the blade starts with a straight line extending from the peripheral wall of the axle radial outwardly.
It is an object of the invention to provide a fluid displacement pump, which overcomes the disadvantages of the heretofore-known devices and methods of this general type and which is further improved in terms of efficiency and backflow prevention, and which allows essentially continuous pumping output with negligible backflow.
With the foregoing and other objects in view there is provided, in accordance with the invention, a fluid displacement pump, wherein said blades having a decreasing thickness from said axles to an outer periphery thereof, and a convex rounded surface extending from said axle outward.
In an alternative embodiment of the invention, the blades increase in thickness from the axle outward. Details of the alternative embodiment will emerge from the following description of the figures.
In accordance with an additional feature of the invention, said rounded surface is defined by a radius of curvature in a radial section of said blades, said radius being greater than a diameter of said blades. Preferably, the radius of curvature is approximately three times the diameter of said blades.
With the above and other objects in view there is also provided, in accordance with the invention, a pump, wherein said blades (9a, 9b) having a thickness (H2) and helically rising along said axle (31) with a spacing (L) greater than the thickness of said blades (9a, 9b), and said blades (9a, 9b) having a convex rounded surface extending from said axle (31) outward.
In a preferred embodiment, the ratio of the spacing between the blade turns (the lead minus the blade thickness) to the thickness of the blades lies between 5/4 and 2.
The axles are preferably cylindrical, i.e., their peripheral wall is defined by mutually parallel lines.
In accordance with an added feature of the invention, the rounded surface is defined by a radius of curvature in a radial section of the blades, the radius being greater than a diameter of the blades. In a preferred embodiment, the radius of curvature is approximately three times the diameter of the blades.
In accordance with another feature of the invention, the blade on each of the axles has a helical rise of approximately 7° and the blades are substantially trapezoidal in radial section from the axle to a periphery thereof.
In accordance with again an added feature of the invention, the blade of one helix of the double helix are spaced apart by a distance defined by the blades of the other helix of the double helix.
In accordance with a concomitant feature of the invention, the blades enclose an angle of between approximately 45° and almost 90° with the cylinder axes.
Other features which are considered as characteristic for the invention are set forth in the appended claims.
Although the invention is illustrated and described herein as embodied in a fluid displacement pump with backflow stop, it is nevertheless not intended to be limited to the details shown, since various modifications and structural changes may be made therein within the scope and range of equivalents of the claim.
The construction and method of operation of the invention, however, together with additional objects and advantages thereof will be best understood from the following description of specific embodiments when read in connection with the accompanying drawings.

Fig. 1 is a partial sectional and side-elevational view of a fluid displacement pump according to the invention;
Fig. 2 is a top plan view onto the impeller blades and the housing of Fig. 1;
Fig. 3 is a plan view of the housing;
Fig. 4 is a plan view onto the impeller blades;
Fig. 5 is a side view of two mutually interengaging blade structures;
Fig. 6 is an enlarged view of the detail indicated in Fig. 5;
Fig. 7 is an axial section through the axle and a blade of a preferred embodiment of the invention;
Fig. 8 is a diagrammatic sectional view of an alternative embodiment of the blade structure;
Fig. 9 is a diagrammatic sectional view of a further alternative embodiment of the blade structure;
Fig. 10 is a diagrammatic section view of yet another alternative embodiment of the blade structure;
Fig. 11 is a diagrammatic sectional view of another alternative embodiment of the blade structure;
Fig. 12 is a diagrammatic sectional view of yet another alternative embodiment of the blade structure;
Fig. 13 is a diagrammatic sectional view of an alternative orientation of the blade structure;
Fig. 14 is an elevational view of two equal orientation impeller blades prior to interengagement; and
Fig. 15 is an elevational view thereof, after the two blades have been inserted into one another.

Referring now to the figures of the drawing in detail and first, particularly, to Fig. 1 thereof, there is seen an elevational view of two interengaging impellers with a section outline of the sidewalls of a housing and a diagrammatic view of a drive system. The fluid pump is a double impeller system, with a first impeller 9A driven by a first gear 14A and a second impeller 9B driven by a second gear 14B. The impeller embodiment is a positive displacement system and, at the same time, a back-pressure membrane. As the ribbed impellers rotate, the fluid flow 11 (e.g., air, liquid, hydraulic fluid) is "packaged" into chamber 30 formed between a cylindrical impeller axle 31, a housing wall 20, and a blade 9B. Each impeller has a respective blade 9A and 9B.
Following the helical path of the chamber 30, each chamber formed between the turns of the blade 9B is closed off by the blade 9A of the adjacent impeller structure. Depending on the rotational speed of the impeller system and the size of the chambers 30, the impellers 9A and 9B form a pressure pump with positive displacement towards a high-pressure chamber. The fluid flow 11 is at a lesser pressure than in the high-pressure chamber, located above the housings in Fig. 1. As the blades 9A and 9B of the impeller rotate, various vertically stacked chambers are opened and closed so that it will result in a positive flow from the bottom to the high-pressure side at the top. At the same time, any pulsations and explosions due, for example, to a combustion of fuel in a chamber on the high-pressure side or any other backpressure will be prevented from flowing back past the blades 9A and 9B. In other words, the impeller pump is always closed with regard to a direct backflow of the fluid out from the high-pressure side.
The impellers 9A and 9B may be driven at variable speed. In order to synchronize the blades 9A and 9B, they are connected via gear wheels 14A and 14B, respectively, connected to their axles 31. A drive 26 is diagrammatically illustrated towards the left of the gear 14A. The drive 26 may be, for example, a gear of a toothed rack, an electrical motor, a feedback system driven by the output of the axles 31, or any similar controlled drive. Any type of speed control may be implemented for the impeller system. It is also possible, of course, the drive the shafts 31 directly with direct drive motors. The two spindles are engaged with the meshing gear wheels 14A and 14B.
Fig. 2 is an axial plan view of the impeller system showing the engagement or meshing of the two blades 9A and 9B and the tight placement of the impeller blades inside the walls 20. The positive displacement force of the impeller system is thus only slightly impaired by backflow and leakage between the impeller blades 9A, 9B and the walls 20 and, negligibly, between the axle 31 and the adjacent blade 9A or 9B. The blades 9A and 9B seal tightly against the housing wall 20. In an exemplary embodiment of the novel fluid pump, the spacing between the outer periphery of the blades and the inner surface of the wall is in the range of a few 0,0254 millimeters (mils), for example 0.1 - 0.4 mm. Depending on its use, the fluid pump may be additionally sealed with a silicon sealing layer provided on the inside of the housing wall and/or on the periphery of the blades 9A and 9B.
With reference to Figs. 2 and 3, the housing of the positive displacement system is defined by walls 20 with rotationally symmetrical portions. In the illustrated embodiment with the two interengaging impellers, the housing has two intersecting circular arches that essentially correspond to the periphery of the blades 9A and 9B in their engagement position. A width D of the housing opening in which the impeller spindles are rotatably disposed corresponds to a sum of the diameters of the impeller blades 9A, 9B minus the overlap O. The overlap O, in turn, corresponds essentially to the rifling depth of the impellers, i.e., the difference in the radius of the blades 9A, 9B and the radius of the shaft 31. The width D may also be expressed as the sum of two times the diameter d of the shaft 31 plus two times the rifling depth of the impellers.
As seen in Figs. 4 and 5, the blades or helical rifling of the blades is offset by approximately 180° so as to distribute the pumping discharge of each of the chambers 30 into the high-pressure side. In other words, it is advantageous for the chambers 30 to reach the top position at which they empty into the high-pressure side alternatingly. In the case of two blades, the offset should thereby be in the neighborhood of 180°.
If three or more impeller spindles are used, the housing 20 requires a corresponding modification and, advantageously, the rotary offset of the impeller rifling may be distributed accordingly by 360°/n, where n is the number of impeller spindles.
The volume of the chambers 30 and the rotational speed of the impellers defines the pump pressure and the volume displacement per time of the impeller injection. With reference to Fig. 6, the volume of each chamber 30 corresponds approximately to the double integral of the differential rotary angle dθ taken through 360° and the differential radius dr taken from the radius r of the shaft 30 to the radius R of the impeller blade 9A, 9B, multiplied with the blade spacing z, minus the volume portion of the adjacent blade that engages into the space in the center between the two spindles.
In order to maximize the seal between the blades, and thus the seal of the backflow-preventing wall, the blades 9A and 9B are modified in terms of their curvature. In that regard, the illustration in Figs. 1, 5, and 6 is simplified to show the blades with a constant thickness from the axle 31 to their peripheries. With reference to Fig. 7, which is a sectional view taken diagonally through the center of the axle 31 of one of the impellers, the blades are curved from the axle outward with regard to their thickness. The measurements and relationships among the various dimensions are best illustrated with reference to a specific example.
In the exemplary embodiment, the blades 9 have a diameter D = 125 mm (5 in). The axle 31 has a diameter d = 25 mm (1 in). The radius r of the blades, therefore, is r = 50 mm (2 in), measured from the periphery of the axle 31 to their outer periphery. The rise angle of the helically winding blades 9 is about 7°. As an intermediate production step, the blades may be tapered by a taper angle ϕ = 3°. That is, the angle α formed between the peripheral wall of the axle 31 and the blade 9 is α = 90° + ϕ = 93° at the top and at the bottom. Furthermore, the blades 9 are curved from the inside out with a radius of curvature R = 400 mm (16 in). The position of the origin of the radius R (i.e., the center of the arc) is defined by the angle ϕ. For instance, if ϕ = 0, then the blades are not tapered, and the origin of R lies on the peripheral wall of the axle 31. If the blades are tapered with ϕ > 0, then the origin of R is moved into the axle 31 by the appropriate amount defined by the angle ϕ. By modeling the novel shape of the blades, the inventor has been able to confirm that a proper and superior seal is created between the interengaging impellers.
Fig. 8 illustrates an alternative in which the blades 9 are only tapered with the angle ϕ. The surfaces are not rounded. In a preferred embodiment of this alternative, the angle ϕ = 3°.
Fig. 9 illustrates yet another alternative. Here, the blades are not tapered, but only curved. Again, the radius R = 400 mm (16 in) and the origin of the arc lies on the peripheral wall of the axle 31. Accordingly, the intersection angle α between the blade 9 and the axle 31 is α = 90°.
Fig. 10 illustrates a further variation. Here, the inventor recognized that certain fluids (usually lower viscosity fluids) require a less proper seal between the blades. Accordingly, here, a spacing L between the blade windings which defines the lead of the impeller, is greater than a thickness H of the blade 9 (note that the distance L is not the lead of the helical winding, the lead would be defined by the spacing L plus the height of the blade, i.e., L+H). Here, the difference is ΔD = (L-H)/2. The reduction from the spacing L to the thickness H may be from 80% to as much as 50%. In other words, a ratio L/H may range from 5/4 to 2. In the embodiments with the blade taper and/or the curvature defined by the radius R, the parameters L and H must be defined in dependence on the distance r from the axle 31. That is, in that case, ΔD = L(r) - H(r) and the spacing L and the height H of the blade 9 is preferably chosen such that ΔD is constant.
Figs. 11 and 12 illustrate yet a further variation of the inventive concept. In Fig. 11, the blades 9 have a bulge in section. That is, the height H of the blade varies from H₁ at the axle 31 to H₂ at approximately half its radial extent, and then returns to the height H₁ at its outer periphery. The embodiment of Fig. 12 is similar, except the blade 9 thins considerably at its outer periphery, to a height H₃ < H₁ < H₂.
The embodiment illustrated in Fig. 13 provides for an attack angle θ between the blade 9 and the axle which is different from 90°. In a preferred embodiment, the angle θ = 70°. It should be understood that the embodiment with the non-orthogonal orientation of the blades, i.e., the angle θ ≠ 90°, is not exclusive of the rounded and/or tapered variations that are illustrated in Fig. 8, 9, 11 and 12. Further, the increased spacing ΔD illustrated in Fig. 10 may be utilized in this embodiment as well.
It will be understood that, of a pair of blades, one may be right-wound and the other may be left-wound. In that case, a counter-rotation of the two blades leads to a rise of both of the spaces 30. If the two blades are wound in the same sense, then the blades will be rotated in the same direction. In the former case, however, a substantially reduced amount of friction will result between the two sets of blades. Also, if the adjacent blades rise in the same sense, the axes must be offset from parallel by twice their lead angle. This illustrated diagrammatically in Figs. 14 and 15.

Claims

A fluid displacement pump, comprising a housing formed with a chamber (30) having a wall (20) defined by two mutually intersecting cylindrical openings defining respective cylinder axes and two axles (31) respectively disposed at and rotatably mounted about respective axes coaxial with said cylinder axes, said axles (31) each carrying a helically rising blade (9a, 9b) sealing against said wall (20) of said housing and engaging into one another, characterized in that said blades (9a, 9b) having a decreasing thickness from said axles (31) to an outer periphery thereof, and a convex rounded surface extending from said axle (31) outward.
The pump according to claim 1, wherein said convex rounded surface is defined by a radius (R) of curvature in a radial section of said blades (9a, 9b), said radius (R) being greater than a diameter (D) of said blades (9a, 9b).
The pump according to claim 2, wherein said radius (R) of curvature is approximately three times the diameter (D) of said blades (9a, 9b).
The pump according to claim 1, wherein said blade (9a, 9b) on each of said axles (31) has a helical rise of approximately 7° and said blades (9a, 9b) are substantially trapezoidal in radial section from said axle (31) to a periphery thereof.
The pump according to claim 1, wherein said blades (9a, 9b) are formed such that a counter-rotation of two interengaging blades (9a, 9b) results in a rising displacement of said blades (9a, 9b).
The pump according to claim 1, wherein said axles (31) are cylindrical axles.
The pump according to claim 1, wherein said blade (9a, 9b) of one helix of said double helix are spaced apart by a distance (z) defined by said blades (9a, 9b) of the other helix of said double helix.
The pump according to claim 1, wherein said cylinder axes and said axles (31) are parallel to one another.
The pump according to claim 1, wherein said axles (31) enclose an angle with one another, and said angle corresponds to twice a rise angle of said blades (9a, 9b).
The pump according to claim 1, wherein said blades (9a, 9b) enclose an angle of between approximately 45° and almost 90° with said cylinder axes.
A fluid displacement pump, comprising a housing formed with a chamber (30) having a wall (20) defined by two mutually intersecting cylindrical openings defining respective cylinder axes and two axles (31) respectively disposed at and rotatably mounted about respective axes coaxial with said cylinder axes, said axles (31) each carrying a helically rising blade (9a, 9b) sealing against said wall (20) of said housing and engaging into one another characterized in that said blades (9a, 9b) having a thickness (H2) and helically rising along said axle (31) with a spacing (L) greater than the thickness of said blades (9a, 9b), and said blades (9a, 9b) having a convex rounded surface extending from said axle (31) outward.
The pump according to claim 11, wherein a ratio of the spacing (L) to the thickness (H2) of the blades (9a, 9b) lies between 5/4 and 2.
The pump according to claim 11, wherein said rounded surface is defined by a radius (R) of curvature in a radial section of said blades (9a, 9b), said radius (R) being greater than a diameter (D) of said blades (9a, 9b).
The pump according to claim 13, wherein said radius (R) of curvature is approximately three times the diameter (D) of said blades (9a, 9b).
The pump according to claim 11, wherein said blades (9a, 9b) have an increasing thickness from said axles (31) radially outward.
The pump according to claim 11, wherein the thickness (H1) of said blades (9a, 9b) at said axle (31) is substantially equal to the thickness (H1) at the outer periphery.
The pump according to claim 11, wherein the thickness (H3) of said blades (9a, 9b) at the outer periphery is smaller than the thickness (H1) at said axle (31).