EP1007822A1 - Moteur rotatif et procede permettant de determiner les contours de ses surfaces de contact - Google Patents

Moteur rotatif et procede permettant de determiner les contours de ses surfaces de contact

Info

Publication number
EP1007822A1
EP1007822A1 EP99925884A EP99925884A EP1007822A1 EP 1007822 A1 EP1007822 A1 EP 1007822A1 EP 99925884 A EP99925884 A EP 99925884A EP 99925884 A EP99925884 A EP 99925884A EP 1007822 A1 EP1007822 A1 EP 1007822A1
Authority
EP
European Patent Office
Prior art keywords
rotor
rotors
gap
axis
contoured
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Withdrawn
Application number
EP99925884A
Other languages
German (de)
English (en)
Inventor
James B. Klassen
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Outland Technologies(USA) Inc
Original Assignee
Outland Technologies(USA) Inc
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Outland Technologies(USA) Inc filed Critical Outland Technologies(USA) Inc
Publication of EP1007822A1 publication Critical patent/EP1007822A1/fr
Withdrawn legal-status Critical Current

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01C—ROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
    • F01C3/00—Rotary-piston machines or engines with non-parallel axes of movement of co-operating members
    • F01C3/06—Rotary-piston machines or engines with non-parallel axes of movement of co-operating members the axes being arranged otherwise than at an angle of 90 degrees
    • F01C3/08—Rotary-piston machines or engines with non-parallel axes of movement of co-operating members the axes being arranged otherwise than at an angle of 90 degrees of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F01C3/085—Rotary-piston machines or engines with non-parallel axes of movement of co-operating members the axes being arranged otherwise than at an angle of 90 degrees of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing the axes of cooperating members being on the same plane

Definitions

  • the present invention relates to rotary positive displacement engines and to methods for determining engagement surface contours for use in the making of rotary positive displacement engines.
  • This invention concerns an advanced rotary positive displacement engine having high power to mass ratio and low production cost.
  • engine as used in this patent document is taken to be a device that converts one form of energy into another. Hence, the term includes both devices which impart energy to the fluid flow (e.g. a pump) and those which employ the fluid flow to generate an energy output (e.g. an external combustion engine for providing a power source) .
  • the reciprocating piston type is most widely used for its low cost of production and efficient sealing, while the turbine has shown that an external combustion engine may offer greater power, partially from high speed.
  • Rotary engines such as the ankel engine have shown higher power-to-weight ratios than reciprocating engines but at the expense of increased fuel consumption.
  • the present invention is a rotary device that offers many of the advantages of these prior art devices without many of their shortcomings .
  • Pumps of the positive displacement type are typically reciprocating or rotary.
  • Many previous rotary combustion engine designs in turn, have been of the single plane type in which rotary motion occurs about axes that are parallel to each other.
  • CvR Engine positive displacement engine which has been developed by Applicant (referred to from time to time herein as a "CvR Engine") and, which is disclosed and issued U.S. Patent No. 5,755,196, the entirety of which is hereby incorporated by reference herein.
  • the present invention provides several significant improvements and advances which are applicable to the CvR engine design which is disclosed in US 5,755,196.
  • Compressible or deformable materials and coatings can provide increased seal performance if they are designed to interfere with the mating surface on the opposite rotor. This can be accomplished by coating a harder material having a negative SSG to bring the surface back to a reduced negative SSG or a positive SSG.
  • Fluid film bearings are used in industry to replace ball bearings or plain bearings in many applications. Fluid films for bearings range from several ten thousandths of an inch to several thousandths of an inch. Having a fluid film between the sealing surfaces of the engine rotors will decrease friction and wear, however, establishing this fluid film requires a correctly designed surface interface. If the surface interface has a gap space which does not account for the other variables which affect the fluid film, extra friction and wear, as well as volumetric efficiency compromises, may result.
  • the present invention is of the rotary positive displacement type, but is in a class by itself.
  • This rotary positive displacement device is believed to be the first rotary engine in which the axes of the moving parts are offset from each other and the moving parts rotate at a constant velocity relative to each other when they are rotating at a constant velocity relative to the casing.
  • the engine is formed by a pair of facing rotors that are axially offset from another and whose faces define chambers that change volume with rotation of the rotors.
  • An engine of this type defines a new class of engines, and includes a minimum number of moving parts, namely as few as two in total .
  • a pump in one aspect of the invention, includes a pair of rotors, both housed on and preferably within the same housing.
  • the housing has an interior cavity having a center.
  • Each rotor is mounted on an axis that passes through the center of the cavity, the respective axes of the rotors being at an angle to each other, with the center of each rotor being at the center of the cavity.
  • the rotors interlock with each other to define chambers .
  • Vanes defined by a contact face on one side of the vane and a side face on the other side of the vane protrude from the rotors.
  • the contact faces of the rotors are defined so that there is constant linear contact between opposing vanes on the two rotors as they rotate.
  • the side faces are preferably concave and extend from an inner end of one contact fact to the outer end of an adjacent contact face, equivalent to the tip of a vane.
  • the side faces and contact faces define walls of chambers that change volume as the rotors rotate.
  • Ports for intake and exhaust are preferably configured to have shapes complementary to the intersecting vanes of the rotors .
  • a method for determining a precise, controllable gap between the sealing surfaces on the rotors. These methods include both mathematical and geometric processes, a well as methods for verifying that the correct contours have been imparted to the surfaces .
  • the vanes on the rotors are provided with mirror-image contoured sealing surfaces which both maintain the desired gap during operation by reducing back-lash, and which also permit efficient reverse operation of the engine .
  • FIG. 1 is an isometric view of a master rotor and slave rotor housed within a ported housing according to one aspect of the invention
  • FIG. 2 is a schematic view showing the interior of the housing of FIG. 1;
  • FIG. 3 is an end view, partially in section, of the housing of FIG. 1;
  • FIG. 4 is a schematic partially in section, of the housing of FIG. 1 showing a cantilevered slave rotor shaft;
  • FIG. 5 is a perspective view of an engine in accordance with a further embodiment of the invention, with the casing the pump being shown separated to expose the internal components thereof, this embodiment of the invention having vanes with mirrored contact surfaces which maintain closer operating tolerances between the vanes and also permit the engine to operate in a reverse direc ion;
  • FIG. 6 is an elevational view of a first half of the engine casing of FIG. 5, showing the port, seal and bushing structure thereof in greater detail;
  • FIG. 7A is a side elevational view of the slave and master rotor of the engine of FIG. 5, showing the engagement of the contact surfaces and the incidence angle between the two rotors ;
  • FIG. 7B is a top plan view of the master and slave rotors of FIG. 7A, showing one of the chambers at its point of maximum volume, containing the fluid received from the inlet port ;
  • FIG. 7C is a bottom plan view of the master and slave rotors of FIG. 7A, showing the chamber at its point of minimum volume, with the bulk of the fluid therefrom having passed to the discharge port;
  • FIGS. 8A-8E are a series of geometric figures showing axes, distances, angles, vectors, and other values associated with the mathematical determination of the contact surface contours in accordance with the present invention.
  • FIGS. 9A-9D are a series views of a visual model illustrating the method by which the contours of the contact surfaces are determined in the present invention, by conceptual rotation of predetermined system axes based on a predetermined mathematical relationship;
  • FIGS. 10A-10D are a series of computer-generated graphical images, illustrating the manner in which the contours of the contact surfaces are determined using the mathematical relationship in accordance with the present invention
  • FIG. 10E is a perspective view of one of the rotors in accordance with the present invention, with the dotted line image showing the area of the contact surface having the contour which is generated through the steps shown in FIGS . 10A-10D.
  • FIG. 11A is a geometric figure, similar to FIG. 8C, showing a revised calculation of the contact surface contours to provide a modified tip-radius form having a slightly flattened shape for enhanced wear characteristics;
  • FIG. 11B is a partial, cross-sectional view of the tip portion of a contact surface contour formed in accordance with the relationship shown in FIG. 11A;
  • FIG. 12 is a schematic view showing the relationship of mirrored contact surfaces, somewhat similar to those shown in FIGS. 7A-7C, with these being configured to provide a predetermined spacing between adjacent contact surfaces so as to provide a predetermined fluid film thickness during operation and also to permit reverse operation of the engine ;
  • FIG. 13A is a partial, enlarged view of adjacent tip portions of the mirrored contact surfaces of FIG. 12, showing the spacing between the tip surfaces in greater detail ;
  • FIG. 13B is a geometric diagram, similar to FIGS. 8C and 11A, illustrating the mathematical determination of the contact surfaces having the clearances which are shown in FIG. 13A;
  • FIG. 14 is an elevational, somewhat diagrammatic view illustrating the determination of the engagement surfaces in accordance with a geometric method which corresponds to the mathematical processes illustrated in FIGS. 8A-13B, in which the gap between the sealing surfaces is controlled by the amount of offset between the apex of a hypothetical cone and the intersection of the axes of the rotors upon which the surfaces are formed;
  • FIGS. 15A - 15E are a series of perspective, somewhat schematic views illustrating the manner in which the contoured contact surfaces on the rotor are formed in accordance with the method of FIG. 14, with the movements of the hypothetical cone corresponding somewhat to those of a tool for machining the surfaces;
  • FIGS. 16A - 16B are perspective, somewhat schematic views showing a first rotor, formed as shown in FIGS. 15A - 15E, in predetermined angular engagement with a second rotor having corresponding engagement surfaces, showing the sealing surface gap which is formed by the offset between the two sets of surfaces;
  • FIG. 17 is a schematic, end view of adjacent sealing surfaces such as those which are shown in FIGS. 16A - 16B, illustrating the manner in which the gap between the sealing surfaces is increased or decreased by rotation of the rotor relative to the hypothetical cone which is shown in FIGS. 14
  • FIG. 18 shows a series of schematic views similar to
  • FIG. 17 showing the different forms of parallel and angular interfacial gaps which can be formed between the sealing surfaces by adjusting variable factors in the methods which are illustrated in FIGS. 14 - 15E;
  • FIGS. 19A - 19C are a series of perspective, somewhat schematic views of a rotor assembly in accordance with an embodiment of the present invention in which relief areas are formed in the sides of the sealing surfaces between the upper and lower ends thereof so as to reduce wear and provide enhanced characteristics for certain applications; and
  • FIG. 20 is a chart demonstrating the relationship between the relative sliding velocity of the sealing surfaces of an engine in accordance with the present invention, as a function of shaft velocity.
  • top and bottom Points on a line bisecting the larger angle formed between offset intersecting axes A and B in the plane defined by axes A and B will be referred to as being at the "top”, while points on the extension of that line bisecting the acute angle between axes A and B will be referred to as being at the "bottom” .
  • FIG. 1 there is shown an engine 10 in accordance with one embodiment of the invention, formed by a housing 12 having an interior surface 14 defining at least a partially spherical cavity, with a central point at the center of bearing 16.
  • a master rotor 20 is mounted for rotation on and within the housing 12 about a first axis A.
  • the master rotor 20 includes a shaft 22 extending along the axis A and has contoured faces 24, 26 forming plural vanes 25 on the other side of the master rotor 20 from the shaft 22.
  • a slave rotor 30 is mounted for rotation on and within the housing 12 about a second axis B.
  • the slave rotor 30 includes a shaft 32 and has contoured faces 34, 36 forming plural vanes 35a on the other side of the slave rotor 30 from the shaft 32.
  • Each of the rotors 20, 30 defines at least part of a sphere, and share a common center coinciding with the center of the cavity.
  • the vanes 25, 35 of the opposed faces of the rotors 20, 30 interlock with each other to define chambers.
  • Axis A and axis B are non-collinear, being at an angle to each other, and intersect at the center of the cavity defined by the housing.
  • the shaft 32 is journalled on an axle 33 (FIG. 9) in this example
  • the shaft 32 may also be cantilevered in the same manner as the shaft 22.
  • the master rotor 20 and slave rotor 30 face each other within the housing in an axial direction, each being predominantly on one side of the common center of the rotors .
  • the portion of the interior surface 14 that is spherical is the portion in which both the vanes of the master rotor 20 and salve rotor 30 rotate. In an extreme position, where the vanes of one rotor extend into the shaft of the other rotor the vanes of both rotors extend into the shafts 22, 32.
  • the shafts 22, 32 are not spherical, but rotationally symmetric.
  • the master rotor 20 and slave rotor 30 should be generally spherical in the portions in which they overlap during operation.
  • the remainder of the rotors 20, 30 and the interior surface 14 need only have rotational symmetry to the extent required to have the rotors 20, 30 rotate in the housing 12.
  • the contoured faces 24, 26, 34, 36 of the mater rotor 20 and slave rotor 30 cooperate with each other and the interior surface 14 of the housing 12 to form chambers 40 (the space between the faces of the rotors) that change volume with rotation of the rotors 20, 30 about the axes A and B respectively.
  • Ports 42 are provided in the housing 12 to allow fluid flow in and out of the chambers.
  • Each contoured face is formed of a contact face 24, 34 and a side face 26, 36 defining vanes (blades) 25, 35 between them.
  • the contact faces 24, 34 form areas of contact between the two rotors 20, 30. Sealing of the chambers 40 is accomplished by close tolerance fit of the rotors 20, 30 against the housing 12 and bearing 16, as well as the relationship of the vanes 25, 35 with respective contact faces 24, 34.
  • the contours of the surfaces in a CvR engine of this type can be determined by defining the contact faces of the rotors by a locus which is formed as the rotors rotate about their respective axes by points on the other rotor, the points of each rotor that define the locus lying along an outer edge of a cone whose central axis is essentially a radius extending outward from the common centers of the rotors at an angle a/2 from a normal to the axis of the other rotor.
  • the contours of the contact surfaces are preferably determined using the methods which are described below.
  • Side faces 26 connect inner ends 27 of one contact face 24 with the outer ends 29 of adjacent contact faces.
  • the side faces 26, unlike the contact faces 24, have a somewhat arbitrary shape. Clearly, they should not stick out beyond the tips 28 of the vanes 25, else they will crash into the side faces 36 of the salve rotor 30.
  • the shape of the side faces 26 can be adjusted for different volumetric ratio changes of the chamber 40 defined between the rotors 20, 30.
  • the chambers 40 may compress to one seventh their maximum size (compression ratio 7:1) in a three vane case. For the embodiment shown by the doted line in FIG. 1 the ratio will be smaller.
  • the point of maximum compression occurs when the vanes 25a, 35a are equidistant from the bottom of their rotation, that is from the line bisecting the acute angle between axes A and B.
  • Enlargement of the chambers 40 may be accomplished by removing material from the side faces 26, 36 to render them concave. Dotted lines F in FIG. 1 show preferred cutting lines. The resulting chambers have considerable volume for the efficient pumping of fluid due to reduction in fluid velocity at the intake and exhaust chambers.
  • the master rotor 20 and salve rotor 30 could conceivably rotate cantilevered on their shafts 22, 32 respectively without additional bearings.
  • contact problems and fluid loss at the center of the cavity poses considerable difficulties.
  • a spherical bearing housing be formed by removal of a partial sphere of material from the center of each of the master rotor 20 and salve rotor; the spherical bearing housing houses bearing 16.
  • the material of the rotors housing the bearing 16 is in fact concave over greater than 180°, creating difficulties in construction.
  • the bearing may be made integral with or otherwise fixed to either rotor, preferably the master rotor 20.
  • the bearing 16 can be loosely fitted in a less than 180 bearing housing, resulting in a greater leakage path, or the bearing may be press fitted into the housing, thermally contracted and inserted into the bearing housing, or slotted for insertion and rotated once inside the bearing housing to present a round bearing surface to the slave rotor.
  • the master rotor 20 is driven by a power source (not shown) through shaft 22. Vanes 25 of rotor 20 push on contact faces 34 of rotor 30 on the side shown on the other side (not shown) contact face 24 of rotor 20 push on vanes 35 of rotor 30.
  • FIGS. 2, 3 and 4 The internal and external configuration of the housing is shown in FIGS. 2, 3 and 4.
  • the housing 12 is preferably formed of two halves 12a and 12b bolted together with bolts 54.
  • the ports 42 are located at opposed sides of the housing, with an intake port 42a and outlet port 42b. Areas 55 show contact areas of vane on contact faces between the master and slave rotors 20, 30. Fluid enters the intake port 42a in expanding chamber 40a. Chamber 40c is at maximum expansion in this rotational position. Chamber 40b is contracting and therefore forces fluid out of port 42b.
  • Chamber 40d is at maximum compression in this rotational position.
  • the ports 42 have peripheries that match the chamber configurations at the point the chambers cross the boundaries of the ports so that as many points as possible of the chamber edge, defined by a pair of vanes 24, 34, cross the port edges at the same time.
  • the trailing edge of the set of vanes beginning to cross the exhaust port or intake port defines the preferred shape of the port at that position.
  • the leading edge of the vanes exiting the intake port or exhaust port defines the preferred shape of the port at that position.
  • the mater rotor is driven by a power source. Rotation of the master and slave rotors with each other causes the chambers 40 to contract while moving from the point of maximum separation of the rotors at the top to the point of minimum separation of the rotors at the bottom. On the other side, the chambers expand. While expanding, the chambers intake fluid, and while contracting the chambers expel fluid, increasing the velocity and/or pressure of the fluid, and energy of the fluid. Thus, energy of the motor driving the pump is converted to energy imparted to the fluid.
  • the parts described here may be made of any suitable materials including plastics and metal, depending on the intended use. Steel may be used for the master rotor 20, while brass may be used for the slave rotor 30. At 10,000 rpm. , a steel and bronze pump is believed to be able to produce 10 hp per lb weight of pump, and 20 hp per lb weight of pump for titanium rotors. As will be described below, care must be taken to provide close tolerance fits of the vanes so that little fluid can escape past the vane contacts and between the rotor and the casing. Material may also be added to the vanes to allow wear.
  • This invention provides a positive displacement rotary pump with high efficiency, believed to be over 90% overall efficiency, and for a pump with eight inches outside diameter, with seven inch diameter rotors, is believed to be able to pump one liter per revolution. 100% rotary motion provides low stress on parts and low vibration. Applications include irrigation, fire fighting, down-hole water and oil pumping, hydraulics, product transfer pumps and high rise building water pumps. b. Mirrored Contact Surfaces.
  • FIGS. 5-7C A preferred embodiment of the invention is shown in FIGS. 5-7C, the engine in this exemplary embodiment being configured for use as a pump, although again it will be understood that the engine can be configured as an external combustion engine or other power source.
  • a particular enhancement featured in the embodiment which is shown in FIGS. 5-7c lies in the mirrored contact surfaces which are provided on the leading and trailing sides of the "lobes".
  • the engine 100 as shown in FIG. 5 includes a master or power rotor 112 which rotates about a first axis A and slave or passive rotor 114 which rotates about a second axis B which is offset from the axis A by an angle ⁇ (see FIG. 7A) .
  • the rotors are housed between the two halves 116a, 116b of an external casing which seals and supports the assembly and also has inlet and outlet ports for the flow of fluid through the engine.
  • Each rotor 112, 114 is partially spherical with a common center, and the casing includes a corresponding spherical cavity 118 which receives and holds the rotors in engagement.
  • the end shafts 120, 122 of the master and slave rotors are supported by the casing.
  • the end 124 of the latter terminates and is fully enclosed within the casing 116, which provides the advantages of simplified sealing and reduced cost of manufacture, although it will be understood that in some embodiments the slave rotor shaft may extend through the exterior of the casing.
  • the master, r ⁇ tor end shaft 120 in turn, extends outwardly from the casing and is connected to a suitable external power source (not shown) , such as an electric, hydraulic or other motor.
  • Each end shaft is supported in a pair of bearings 126 and 128 to maintain shaft stability and eliminate end play.
  • the inner bearings 126 include conical bearing faces (not shown) which engage corresponding conical tapers 129a, 129b on the backs of the rotors, so as to react against thrust loads and maintain the rotors in proper engagement .
  • the bearings are received in corresponding cavities 130, 132, with lubricant being supplied to the cavities through a series of ports 134.
  • the bearings are preferably high speed fluid film bushings, i.e., bushings which run on a thin film of air, oil, water, etc., although it will be understood that other forms of high speed bearings may be employed in some embodiments.
  • a continuous elastomeric seal 136 is retained in a channel 138 which extends completely around the rotor chamber and shafts, and includes a ring seal 140 which surrounds the master rotor and shaft where this exits the housing; the seal 136 may suitably be formed of a moldable polyurethane material .
  • the clamping force of the two casing halves against the elastomeric member provides the low pressure seal for the assembly, while the fluid pressure acting outwardly against the elastomeric material creates the high pressure seal .
  • the casing also includes an inlet port 142 and an outlet port 144, which communicate with the rotor chamber 118 and via which the fluid enters and leaves the engine; the inner edges 146, 148 of the ports, where these meet the spherical rotor chamber, have a shape which matches the corresponding edges of the contact surfaces define the sealed chamber between the rotors (which shape will be described in greater detail below) , while the outer edges 150, 152 of the ports are round for connection to conventional circular cross-section tubing or other conduits .
  • FIG. 7A shows the engagement of the first and second mirrored contact surfaces 160, 162 on each vane 164, and the contact surfaces 166, 168 on the corresponding cavity 170.
  • This engagement forms a substantially sealed chamber which changes in volume with rotation of the rotors.
  • each vane or lobe is provided with two mirror image contact surfaces, i.e., a leading contact surface and a mirror image trailing contact surface.
  • FIG. 7B is the top or "overhead” view of the master and slave rotors 112, 114.
  • the lobes 164a, 164b, etc. of the master rotor 112 are angularly spaced so as to define a plurality of angularly spaced cavities 172, and the lobes 174 on the slave rotor define corresponding cavities 176.
  • each lobe is received in the corresponding cavity in the opposite rotor i.e., the master rotor lobes 164 are received in the cavities 176 in the slave rotor, and the slave rotor lobes 174 are received in the cavities 172 in the master rotor.
  • the area in the center of the rotors, between the lobes on either side, is sealed by a ball 175 or other generally spherical body.
  • each lobe engages the corresponding contact surfaces on each socket (these being the contact surfaces of the lobes on either side of the socket), as indicated at the areas 178. Consequently, a series of sealed chambers 180a, 180b, 180c are formed about the end of the master rotor, between the ends or "heads" of the lobes in the bottom of the cavities, and a corresponding series of sealed chambers 182a, 182b, 182c are thus formed around the end of the slave rotor.
  • the chambers change in volume with rotation of the rotor assembly, in the direction indicated by arrow 184.
  • the volume of the chamber increases as these rotate past the inlet port 142 (see FIG. 5) , thus drawing fluid into the pump.
  • the ports are shaped so that each chamber moves out of register with the inlet port just as the chamber reaches its maximum volume (see chamber 180b in FIB. 7B) , and just before the chamber begins to rotate into register with the outlet port 144.
  • the chambers then decrease in volume as they rotate the outlet port, forcing the fluid outwardly, and reach a minimum volume at the bottom of the cycle (see chamber 182c in FIG.
  • the embodiment having the lobed vane structure with mirrored leading and mirrored contact surfaces has several advantages over the device which is shown in FIGS. 1-4. Firstly, the use of mirrored contact surfaces enables the engine to run and develop pressure in either direction of rotation. This is because the mirrored contact surface lobes do not require the force of the faster (power) rotor vanes pushing against the slave rotor vanes in order to maintain a contact seal. Moreover, the mirrored contact surfaces on the lobes enable these to maintain an acceptable fluid film between the surfaces at a wide range of operating speeds and fluid viscosities . Maintaining a thin fluid film between the contact surfaces is advantageous for reducing wear and friction.
  • the mirrored contact surfaces control the amount of "backlash" between the slave and power rotors, so that only the predetermined amount of rotation is allowed between the two, which in turn defines the maximum clearance/fluid thickness there can be between the leading and trailing contact surfaces of the lobes.
  • the fluid film at the leading contact surface of each lobe will tend to be slightly less than that of the trailing contact surface; however, depending on operating speed, back pressure, fluid viscosity and other factors, an equilibrium level is achieved in which a fluid film exists between both leading and trailing surfaces .
  • Additional advantages include increased strength of the rotor lobes, since the area between the mirrored contact surface (i.e., the backs of the contact surfaces) can be filled in, so that the back side of each of the faces is reinforcing the other, giving the lobes strength comparable to that of a gear tooth. Also, because of the higher strength, it is possible to operate the pump at higher pressures, which is advantageous in increasing the power ratio, or power density, of the pump.
  • FIGS. 8A-8D provide a series of graphical representations of axes, vectors, angles, and other values associated with the mathematical computation of the contact surfaces of the vanes/lobes, as follows:
  • FIG. 8A shows the orientation of the two rotor axes, Axis 1 and Axis 2, intersecting at 0 and placed at an angle A° apart.
  • the line 0-0 is initially in the plane of the two axes and bisects the direction of each, so that it makes an angle of (90+A/2) ° with each axis direction.
  • Point Q is a radial line on the surface of a sphere o radius R, which is a point locating the working surface of the rotor attached to shaft [2] .
  • the plane P formed by the line 0-0 and 0Q will be a plane that changes orientation in space.
  • a set of unit vectors can be used to describe the orientation of plane P in space. As shown in FIG. 8C, let ul be the first vector, directed along 0-0, and is defined in terms of the vector R 0 .
  • the vector uO is a unit vector with direction along 0-0, which changes with rotation, as does ul .
  • the outer edge of the surface determined by Q is shown in FIG. 8D. Also shown is the rotation of plane P for different rotations of the shafts.
  • the total angular twist S, along the axis 0-0 in any general position can be most easily obtained by determining the angular change in the normal to the plane P, which is the unit vector u2. This vector is always directed along the tangent to the path of Q or Q', and has already been defined.
  • the untwisted position of the plane P can be obtained by rotating the plane and its initial normal direction vector u2 about the z axis in the xy plane through the angle ⁇ , to a new position u2' 0 and again about the plane OQ'Q" through the angle ⁇ with the Z axis in the x-y plane through the angle ⁇ , to a new position u2 'o and in the plane °Q'Q" until it makes an angle ⁇ with the Z axis to a final position u2" 0 .
  • FIGS. 9A-9D are a series of views of a model which provides a visual representation of the relationships between axes and points in the system described by the mathematical process above.
  • FIG. 9A shows the "start" position, in which the axes 1 and 2 correspond in angular relationship to the axes of the master and slave rotors, the length 0-Q represents the radius of the rotor, and the point Q, on a line normal to axis R, represents one point along the contact surface of the lobe.
  • the offset between 0-0 and Q represents the surface depth of the lobe.
  • FIGS. 9A-9D show rotation of Axis 1 90° from the start position to the final position; it will be understood, however, that determination of the line is ordinarily carried out in small degree increments, so as to define a smooth, continuous contour.
  • FIG. 9B shows the model 190 with the Axes 1 and 2 having been rotated together by an angle ⁇ of 90°, so that axis R swings from the vertical alignment (for purposes of illustration) shown in FIG. 9A to the horizontal alignment in FIG. 9B .
  • Axis 1 held stationary, Axis 2 is rotated back by an angle - ⁇ , which is equal to ⁇ but in the reverse direction, rotating axis R to the position which is shown in FIG. 9C.
  • the axis R is rotated by the amount ⁇ g which is calculated in accordance with the mathematical system described above, bringing point Q to its final position Q", as shown in FIG. 9D.
  • FIG. 9D also includes a broken- line image 192 which shows the original position of point Q at the start point shown in FIG. 9A.
  • FIGS. 10A-10D are a series of views similar conceptually to FIGS. 9A-9D, but showing the manner in which the above process is used to generate or determine a contoured line 194 in a computer plotting program.
  • the point Q is moved sequentially from position to position line 194, with each rotation of the Axes 1 and 2.
  • Connecting the dots i.e., the position of point Q at each position of Axis 1, a continuous contour line is created which corresponds to the contour line along one of the contact surfaces, such as the contact surface 160 on lobe 164, as shown in FIG. 10E.
  • the offset establishes sufficient clearance between the contact surfaces to establish the fluid film and avoid the parts rubbing directly on one another.
  • the amount of the offset is determined on a basis of fluid type and viscoscity, operating speeds and pressures, and materials characteristics, along with other factors.
  • a "negative" offset may be used, so as to cause some interference between the contact surfaces which forms an enhanced seal; this may be particularly desirable for high-pressure, low-speed applications .
  • the three-dimensional surface is generated by one of two methods. Firstly, the contour line can simply be scaled down towards the center of the rotor, in which case the clearances and thickness of the fluid film will also decrease towards the center accordingly.
  • the contour line can be recalculated at the smallest radius at the lobes/vanes, with the intermediate contour lines defined accordingly, so as to give a constant gap/fluid film thickness across the entire contact surface; this approach may be particularly advantageous where the fluid contains particulates of a known size, and it is therefore important to maintain a fluid film which is thick enough to hold these particulates without them being forced into the contact surfaces. Whichever approach is used, one contour can be calculated for the leaving contact surface of the lobe and then reversed for the mirror image trailing contact surface, or vice versa.
  • FIGS. 10A-10D illustrate the manner in which these calculations are employed to produce a computer generated plot of the contour lines
  • FIG. 10E is a partial perspective view of one of the rotors, showing the position of the contour line which has been produced in FIGS. 10A-10D.
  • the working surface would normally follow a radial line towards the center 0, resulting in a film thickness that tapers towards the center.
  • the relative sliding velocity between adjacent lobes will be highest at the outside, so a larger thickness of film there seems reasonable.
  • c is the distance between centers of adjacent tips (measured along the arc of the surface of a sphere of radius R) .
  • the arc length C is the distance between like lobe shapes (circular pitch length) .
  • s is the arc length taken up by the tip
  • t is the film thickness (or net interference)
  • n is an integer number of pitch lengths to make up a full circle. If s and t are chosen,
  • variable spacing As this would help to alleviate the production of a pure tone noise (having a single frequency component) emanating from the running pump. Variable spacing would produce other frequency components, grouped around the running speed frequency and its harmonics (sidebands) . The effect should reduce the overall noise level slightly, but more importantly, be less annoying for personnel in the vicinity.
  • rotor unbalance could be produced for random spacing. If the spacing were arranged symmetrically in pairs, unbalance can be prevented, but the beneficial effect of staggered spacing would be reduced. If the unbalance were the result of a particular arrangement, each rotor could be balanced individually before final assembly. For uniform spacing, whether the number of rotors n is an even or an odd number, balance would be maintained.
  • FIGS . 13-19 illustrate a method for geometric determination of the contact surface contours consistent with the mathematical calculations described above, but which corresponds more directly to an actual manufacturing process for forming the surfaces, as by hobbing material from a blank so as to form the lobes and surfaces .
  • the "lift off clearance” is the thickness of the fluid film between the sealing surfaces of the two rotors when the engine is operating in its intended mode.
  • “Lift off clearance” is affected by the speed of the engine, the viscosity of the fluid medium, and the differential pressure between the inlet port and the discharge port . Contact happens when the one or two or all of these factors is insufficient to maintain a fluid film between the mating surfaces.
  • the contact characteristic describes how the sealing surfaces mate when the fluid film is not sufficient to achieve "lift off".
  • the three basic types of contact are (1) Full radial contact. (2) Inner radial contact. (3) Outer radial contact. These characteristics can be different at different angles of rotor rotation.
  • Maintaining a fluid film is desirable to reduce wear, as well as to allow entrained particles to pass between the sealing surfaces without damaging the particles or causing excessive abrasion to the sealing surface.
  • U.S. 5,755,196 describes a CvR engine configuration with a "contact” or “close tolerance” seal design which does not optimize or account for the "lift off situation".
  • This type of surface geometry relies on a line to line seal between the rotors and is intended to operate with each rotor sliding on the other rotor without consideration of the fluid film between the rotors.
  • the radial difference of the surface speed in this contact zone may make up for the variation in gap thickness when the engine is operating at very low pressures, (relative surface speed is greater at points further from the rotational center) But the fluid film "rigidity" is not linear with the thickness of the film which the surface speed is a linear relationship with the distance from center Ideally then, if surface speed was the only consideration, then the SSG should increase at points further from center, but only enough to establish a consistent fluid film pressure .
  • the fluid film is influenced increasingly by the pressurized fluid which is moving past this area.
  • the fluid film resulting from the surface speed is affected greatly by the distance from center and requires an increasing surface gap towards the outside of the engine.
  • the fluid film resulting from the differential pressure between the output port and the input port is independent of the distance from center and requires a more consistent gap clearance. The more the fluid film is affected by the pressure differential of the fluid, the more consistent the radial gap clearance must be to achieve maximum efficiency and wear characteristics .
  • the present invention provides methods for determining, defining, and/or constructing this more consistent gap clearance, as well as a method for determining, defining, and/or constructing an engine with a gap clearance that also takes into account the surface speed of each rotor on the other to maximize the "liftoff effect" of the fluid film between the rotors .
  • the methods can also be combined to account for other variables including the change in relative surface speed which occurs at different angular rotor positions.
  • contact between the sealing surfaces may occur during start-up under high pressure, but should not continue when the engine is operating in its intended mode.
  • U.S. 5,755,196 describes a surface which is defined by the movement of a cone being rotated around the opposite rotor axis. To achieve the seal surface of the prior art, the apex of the cone should be as close to the center of the spherical center of the rotors as possible.
  • the sealing surface of the present invention can also be described with the movement of a cone around the opposite rotor axis, but the cone of this present invention is positioned intentionally above or below the spherical center of the rotors.
  • a spherical rotor RA is positioned for rotation about its center axis AA.
  • a second axis AB is positioned at an angle X to axis AA.
  • a cone C is positioned with its center axis collinear with a line Y that bisects the obtuse angle between axis AA and axis AB .
  • the cone C is positioned on line Y with its apex X below the point P where the two rotor axes intersect .
  • the apex of the cone must be positioned above the point P. (The smaller the angle of the cone, the more its apex must be positioned off center to achieve a given gap clearance or interference.)
  • the spherical rotor and the cone are then rotated around their respective axes (i.e., cone C rotates on axis AB at a fixed angle thereto) and the path of the cone is removed from the spherical rotor.
  • This will define the "seal surface" S of one side of one vane x on the rotor RA.
  • the rotor is then rotated toward the first cone and another cone shape C is positioned with its axis collinear with the line Y.
  • This cone has the same angle as the first cone and it is positioned with its apex the same distance from center but on the opposite side of point P (see FIG. 15) .
  • This cone is added to the rotor RA and becomes the "seal tip" T of this seal face, as is shown in FIG. 15E.
  • the sequence is then repeated for the second rotor RB (See FIGS.16A-16B) with a cone which is positioned along the center axis of the adjacent "seal tip" T cone of the rotor RA.
  • the engine will have a -predetermined parallel interface gap IG between mating surfaces as is shown most clearly in FIG. 16B.
  • the other gap configuration which can be used on its own or in combination with the "offset cone” gap configuration, is the “angular interfacial gap”.
  • An angular interfacial gap may offer performance benefits for certain applications.
  • the centrifugal force of the rotation of the engine could be used to force particulate matter entrained in the fluid to the periphery of the engine chamber.
  • an angular interfacial gap with a larger gap at the periphery of the rotors would allow the particles to pass through the thicker fluid film, while a more efficient seal could be maintained closer to the center of the rotors where the fluid film is thinner.
  • a characteristic of the "parallel interfacial gap” compared to the “angular interfacial gap” is that the “parallel interfacial gap” method creates a consistent SSG for the entire seal surface.
  • the "angular IG” method (of rotating the seal surface relative to the rest of its rotor) , only changes the gap clearance in a plane that is perpendicular to the rotational axis of the rotor.
  • a reduced gap clearance can be achieved in this area using the Angular IG method or a combination of the Angular IG method and the Parallel IG method of changing the gap at the higher relative speed areas of the seal surface .
  • the surface speed also reduces, but the angular IG method will increase this gap.
  • the cone To increase the gap clearance at some places but not at TDC, it is necessary to sue the Parallel IG method of achieving the desired gap, but the cone must be moved dynamically along its axis as the rotors are rotated during the shaping process .
  • the transitional gap between the rotors changes from an angular interfacial gap to a parallel interfacial gap and on to an angular interfacial gap at an angle in opposition to the initial angular interfacial gap.
  • cone shape described above is the ideal shape, and the simplest to calculate and design, it will be understood that other similar shapes (such as a portion of a much larger cone or simply a sharp edge) could be used, however, as the mating surface is designed to maintain the desired SSG as both rotors spin at the same speed.
  • a contact CMM machine could be used to determine a number of points on the surface of a completed rotor, and establish what the seal surface characteristic is.
  • the most basic way of determining if a rotor design has been manufactured according to the present invention is to create a plane which is perpendicular to a point on the seal face (or seal tip) which passes through the spherical center of the sphere . Two points on the seal face or seal tip surface which are also on this plane will be connected and extended toward the spherical center of the engine .
  • a rotor face with a parallel interfacial gap will result in the extended line passing consistently to the contact surface lobe side of the Spherical center.
  • a rotor face with an angular interfacial gap may result in the extended line passing through the spherical center of the rotors or on either side, depending on the angle, and on the magnitude of the gap.
  • the extended line of an angular interfacial gap will pass through the spherical center or to the side of the spherical center which is away from the mass of the seal surface lobe.
  • a rotor face with a reverse angular interfacial gap will result in the extended line passing consistently on the side of the spherical center which is away from the mass of the seal surface lobe .
  • a rotor face with an interfering parallel interfacial gap will result in the extended line passing consistently on the side of the spherical center which is away from the mass of the seal surface lobe.
  • a rotor face with an interfering angular interfacial gap will also result in the extended line passing consistently on the side of the spherical center which is away from the mass of the seal surface lobe.
  • a rotor face with an interfering reverse angular interfacial gap will result in the extended line passing consistently on the side of the spherical center which is toward the mass of the seal surface lobe .
  • FIGS. 19A-19C illustrate in the embodiment of the present invention in which the sealing surfaces are shaped so as to provide actual fluid sealing during only selected portions of the rotation of the assembly, i.e., at those points during the rotation where the seal is required in order to maintain efficiency.
  • This configuration is advantageous in a number of applications, including for use with pumping sheer sensitive or abrasive fluids, and for enhanced wear characteristics.
  • the sealing surfaces 200 on the vanes 202 of the two rotors 204, 206 are each formed with a recess or channel area 208 which extends radially across the rotor base and separates the sealing surface segments 210, 212 which lie proximate the tip and at base portions of the contoured face.
  • the sealing surface segments 210, 212 are formed in accordance with the methods described above, i.e., these are configured to form the requisite seal with the corresponding segments on the adjoining contoured face, with a predetermined gap as desired. Since the sealing segments are formed at the top and bottom of each surface, the rotors form an effective seal only when the chambers defined thereby are approximately at top and bottom dead center, as is shown in FIGS. 19A and 19B.
  • the channels 208 eliminate direct contact between the two sealing surfaces so as to form a relief gap 220, as is shown in FIG. 19C.
  • the relief gap reduces sheer stresses on fluid in this area, and also allows particulate or abrasive material to pass therethrough without causing wear against the sealing surfaces. Furthermore, the relief gap reduces wear by eliminating a potential content between the sealing surfaces during the intermediate phases of the engine cycle, even in applications not being used with abrasive fluids. Since sealing is only critical when the chambers are at top and bottom dead center, these advantages are achieved without significant cost to the overall efficiency of the engine. It is to be recognized that these and various other alterations, modifications, and/or additions may be introduced into the constructions and arrangements of parts described above without departing from the spirit or ambit of the present invention as defined by the appended claims.

Abstract

Moteur rotatif (10) amélioré et procédé permettant de déterminer les contours de ses surfaces d'étanchéité (24, 34). Ce moteur permet de maintenir, durant la rotation, un espace optimal (IG) prédéterminé entre ses surfaces d'étanchéité. Cet espace peut être parallèle ou oblique, positif ou négatif, de façon à former un contact avec interférence. Les rotors (112, 114) du moteur peuvent présenter des surfaces d'étanchéité en miroir (160, 162, 166, 168), qui empêchent l'apparition d'un jeu et d'un espace mort excessifs et qui permettent un fonctionnement en sens inverse efficace. Les surfaces d'étanchéité (200, 212) peuvent également présenter des évidements (208), qui permettent d'interrompre le contact au niveau de certains points du cycle de rotation, d'améliorer les caractéristiques d'usure et/ou d'adapter le moteur à des fluides abrasifs ou sensibles au cisaillement.
EP99925884A 1998-05-26 1999-05-26 Moteur rotatif et procede permettant de determiner les contours de ses surfaces de contact Withdrawn EP1007822A1 (fr)

Applications Claiming Priority (3)

Application Number Priority Date Filing Date Title
US8683898P 1998-05-26 1998-05-26
US86838P 1998-05-26
PCT/US1999/011642 WO1999061753A1 (fr) 1998-05-26 1999-05-26 Moteur rotatif et procede permettant de determiner les contours de ses surfaces de contact

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EP1007822A1 true EP1007822A1 (fr) 2000-06-14

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AU (1) AU4208599A (fr)
CA (1) CA2300420C (fr)
WO (1) WO1999061753A1 (fr)

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US6705161B1 (en) 2000-08-08 2004-03-16 Outland Technologies (Usa), Inc. Positive displacement flow meter method and apparatus
DE60219441T2 (de) 2001-01-30 2008-03-13 Outland Technologies, Inc. Verdrängervorrichtung, -verfahren und -apparat zur bereitstellung einer minimalkontaktdichtung
CA2440304C (fr) 2001-02-08 2010-05-04 Outland Technologies (Usa), Inc. Dispositif rotatif a deplacement direct
US8602758B2 (en) 2008-09-17 2013-12-10 Exponential Technologies, Inc. Indexed positive displacement rotary motion device
US8562318B1 (en) 2009-08-20 2013-10-22 Exponential Technologies, Inc. Multiphase pump with high compression ratio
WO2019113704A1 (fr) 2017-12-13 2019-06-20 Exponential Technologies, Inc. Dispositif à écoulement de fluide rotatif
US11168683B2 (en) 2019-03-14 2021-11-09 Exponential Technologies, Inc. Pressure balancing system for a fluid pump

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US32372A (en) * 1861-05-21 John jones
US1379653A (en) * 1917-06-13 1921-05-31 Alvin H Shoemaker Rotary engine
GB632462A (en) * 1946-05-17 1949-11-28 Bendix Aviat Corp Improvements in or relating to gear pumps or motors
US3236186A (en) * 1963-04-29 1966-02-22 Wildhaber Ernest Positive-displacement unit
DE2364281A1 (de) * 1973-12-22 1975-06-26 Juergen Schukey Kompressionsmaschine
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CA2300420C (fr) 2005-05-03
WO1999061753A1 (fr) 1999-12-02
CA2300420A1 (fr) 1999-12-02
AU4208599A (en) 1999-12-13

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