Rotary pump with exclusively hydrodynamically suspended impeller

A pump assembly 1, 33, 200 adapted for continuous flow pumping of blood. In a particular form the pump 1, 200 is a centrifugal pump wherein the impeller 100, 204 is entirely sealed within the pump housing 2, 201 and is exclusively hydrodynamically suspended therein as the impeller rotates within the fluid 105 urged by electromagnetic means external to the pump cavity 106, 203. Hydrodynamic suspension is assisted by the impeller 100, 204 having deformities therein such as blades 8 with surfaces tapered from the leading edges 102, 223 to the trailing edges 103, 224 of bottom and top edges 221, 222 thereof.

FIELD OF THE INVENTION
 This invention relates to rotary pumps adapted, but not exclusively, for
 use as artificial hearts or ventricular assist devices and, in particular,
 discloses in preferred forms a seal-less shaft-less pump featuring open or
 closed (shrouded) impeller blades with at least parts of the impeller used
 as hydrodynamic thrust bearings and with electromagnetic torque provided
 by the interaction between magnets embedded in the blades or shroud and a
 rotating current pattern generated in coils fixed relative to the pump
 housing.
 BACKGROUND ART
 This invention relates to the art of continuous or pulsatile flow rotary
 pumps and, in particular, to electrically driven pumps suitable for use
 although not exclusively as an artificial heart or ventricular assist
 device. For permanent implantation in a human patient, such pumps should
 ideally have the following characteristics: no leakage of fluids into or
 from the bloodstream; parts exposed to minimal or no wear; minimum
 residence time of blood in pump to avoid thrombosis (clotting); minimum
 shear stress on blood to avoid blood cell damage such as haemolysis;
 maximum efficiency to maximise battery duration and minimise blood
 heating; and absolute reliability.
 Several of these characteristics are very difficult to meet in a
 conventional pump configuration including a seal, i.e. with an impeller
 mounted on a shaft which penetrates a wall of the pumping cavity, as
 exemplified by the blood pumps referred to in U.S. Pat. No. 3,957,389 to
 Rafferty et al., U.S. Pat. No. 4,625,712 to Wampler, and U.S. Pat. No.
 5,275,580 to Yamazaki. Two main disadvantages of such pumps are firstly
 that the seal needed on the shaft may leak, especially after wear, and
 secondly that the rotor of the motor providing the shaft torque remains to
 be supported, with mechanical bearings such as ball-bearings precluded due
 to wear. Some designs, such as U.S. Pat. No. 4,625,712 to Wampler and U.S.
 Pat. No. 4,908,012 to Moise et al., have overcome these problems
 simultaneously by combining the seal and the bearing into one hydrodynamic
 bearing, but in order to prevent long residence times they have had to
 introduce means to continuously supply a blood-compatible bearing purge
 fluid via a percutaneous tube.
 In seal-less designs, blood is permitted to flow through the gap in the
 motor, which is usually of the brushless DC type, i.e. comprising a rotor
 including permanent magnets and a stator in which an electric current
 pattern is made to rotate synchronously with the rotor. Such designs can
 be classified according to the means by which the rotor is suspended:
 contact bearings, magnetic bearings or hydrodynamic bearings, though some
 designs use two of these means.
 Contact or pivot bearings, as exemplified by U.S. Pat. No. 5,527,159 to
 Bozeman et al. and U.S. Pat. No. 5,399,074 to Nose et al., have potential
 problems due to wear, and cause very high localised heating and shearing
 of the blood, which can cause deposition and denaturation of plasma
 proteins, with the risk of embolisation and bearing seizure.
 Magnetic bearings, as exemplified by U.S. Pat. No. 5,350,283 to Nakazeki et
 al., U.S. Pat. No. 5,326,344 to Bramm et al. and U.S. Pat. No. 4,779,614
 to Moise et al., offer contactless suspension, but require rotor position
 measurement and active control of electric current for stabilisation of
 the position in at least one direction, according to Earnshaw's theorem.
 Position measurement and feedback control introduce significant
 complexity, increasing the failure risk. Power use by the control current
 implies reduced overall efficiency. Furthermore, size, mass, component
 count and cost are all increased.
 U.S. Pat. No. 5,507,629 to Jarvik claims to have found a configuration
 circumventing Earnshaw's Theorem and thus requiring only passive magnetic
 bearings, but this is doubtful and contact axial bearings are included in
 any case. Similarly, passive radial magnetic bearings and a pivot point
 are employed in U.S. Pat. No. 5,443,503 to Yamane.
 Prior to the present invention, pumps employing hydrodynamic suspension,
 such as U.S. Pat. No. 5,211,546 to Isaacson et al. and U.S Pat. No.
 5,324,177 to Golding et al., have used journal bearings, in which radial
 suspension is provided by the fluid motion between two cylinders in
 relative rotation, an inner cylinder lying within and slightly off axis to
 a slightly larger diameter outer cylinder. Axial suspension is provided
 magnetically in U.S. Pat. No. 5,324,177 and by either a contact bearing or
 a hydrodynamic thrust bearing in U.S. Pat. No. 5,211,546.
 A purging flow is needed through the journal bearing, a high shear region,
 in order to remove dissipated heat and to prevent long fluid residence
 time. It would be inefficient to pass all the fluid through the bearing
 gap, of small cross-sectional area, as this would demand an excessive
 pressure drop across the bearing. Instead a leakage path is generally
 provided from the high pressure pump outlet, through the bearings and back
 to the low pressure pump inlet, implying a small reduction in outflow and
 pumping efficiency. U.S. Pat. No. 5,324,177 provides a combination of
 additional means to increase the purge flow, namely helical grooves in one
 of the bearing surfaces, and a small additional set of impellers.
 U.S. Pat. No. 5,211,546 provides 10 embodiments with various locations of
 cylindrical bearing surfaces. One of these embodiments, the third,
 features a single journal bearing and a contact axial bearing.
 Embodiments of the present invention offer a relatively low cost and/or
 relatively low complexity means of suspending the rotor of a seal-less
 blood pump, thereby overcoming or ameliorating the problems of existing
 devices mentioned above.
 SUMMARY OF THE INVENTION
 According to one aspect of the present invention, there is disclosed a
 rotary blood pump for use in a heart assist device or like device, said
 pump having an impeller suspended in use within a pump housing exclusively
 by hydrodynamic thrust forces generated by relative movement of said
 impeller with respect to and within said pump housing.
 Preferably at least one of said impeller or said housing includes at least
 one deformed surface which, in use, moves relative to a facing surface on
 the other of said impeller or said housing thereby to cause a restriction
 in the form of a reducing distance between the surfaces with respect to
 the relative line of movement of said deformed surface thereby to generate
 relative hydrodynamic thrust between said impeller and said housing which
 includes everywhere a localized thrust component substantially and
 everywhere normal to the plane of movement of said deformed surface with
 respect to said facing surface.
 Preferably the combined effect of the localized normal forces generated on
 the surfaces of said impeller is to produce resistive forces against
 movement in three translational and two rotational degrees of freedom thus
 supporting the impeller for rotational movement within said housing
 exclusively by hydrodynamic forces.
 Preferably said thrust forces are generated by blades of said impeller.
 More preferably said thrust forces are generated by edges of said blades of
 said impeller.
 Preferably said edges of said blades are tapered or non-planar so that a
 thrust is created between the edges and the adjacent pump casing during
 relative movement therebetween.
 Preferably said edges of said blades are shaped such that the gap at the
 leading edge of the blade is greater than at the trailing edge and thus
 the fluid which is drawn through the gap experiences a wedge shaped
 restriction which generates a thrust.
 Preferably the pump is of centrifugal type or mixed flow type with impeller
 blades open on both front and back faces of the pump housing.
 Preferably the front face of the housing is made conical, in order that the
 thrust perpendicular to the conical surface has a radial component, which
 provides a radial restoring force to a radial displacement of the impeller
 axis during use.
 Preferably the driving torque of said impeller derives from the magnetic
 interaction between permanent magnets within the blades of the impeller
 and oscillating currents in windings encapsulated in the pump housing.
 Preferably said blades include magnetic material therein, the magnetic
 material encapsulated within a biocompatible shell or coating.
 Preferably said biocompatible shell or coating comprises a diamond coating
 or other coating which can be applied at low temperature.
 Preferably internal walls of said pump which can come into contact with
 said blades during use are coated with a hard material such as titanium
 nitride or diamond coating.
 Preferably said impeller comprises an upper conical shroud having said
 taper or other deformed surface therein and wherein blades of said
 impeller are supported below said shroud.
 Preferably said impeller further includes a lower shroud mounted in opposed
 relationship to said upper conical shroud and whereas said blades are
 supported within said upper and said lower shroud.
 Preferably said deformed surface is located on said impeller.
 Preferably said deformed surface is located within said housing.
 Preferably forces imposed on said impeller in use, other than hydrodynamic
 forces, are controlled by design so that, over a predetermined range of
 operating parameters, said hydrodynamic thrust forces provide sufficient
 thrust to maintain said impeller suspended in use within said pump
 housing.
 Preferably at least one face of the housing is made conical, in order that
 the thrust perpendicular to it has a radial component, which provides a
 radial restoring force to a radial displacement of the impeller axis.
 Similarly, an axial displacement toward either the front or the back face
 increases the thrust from that face and reduces the thrust from the other
 face. Thus the sum of the forces on the impeller due to inertia (within
 limits), gravity and any bulk radial or axial hydrodynamic force on the
 impeller can be countered by a restoring force from the thrust bearings
 after a small displacement of the impeller within the housing relative to
 the housing in either a radial or axial direction.
 In a preferred embodiment, the impeller driving torque derives from the
 magnetic interaction between permanent magnets within the blades of the
 impeller and oscillating currents in windings encapsulated in the pump
 housing.
 In a further broad form of the invention there is provided a rotary blood
 pump having an impeller suspended exclusively hydrodynamically by thrust
 forces generated by the impeller during movement in use of the impeller.
 Preferably said thrust forces are generated by blades of said impeller or
 by deformities therein.
 More preferably said thrust forces are generated by edges of said blades of
 said impeller.
 Preferably said edges of said blades are tapered.
 In an alternative preferred form said pump is of axial type.
 Preferably within a uniform cylindrical section of the pump housing,
 tapered blade edges form a radial hydrodynamic bearing.
 In a further broad form of the invention there is provided a rotary blood
 pump having a housing within which an impeller acts by rotation about an
 axis to cause a pressure differential between an inlet side of a housing
 of said pump and an outlet side of the housing of said pump; said impeller
 suspended exclusively hydrodynamically by thrust forces generated by the
 impeller during movement in use of the impeller.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS
 The pump assemblies according to various preferred embodiments to be
 described below all have particular, although not exclusive, application
 for implantation in a mammalian body so as to at least assist, if not take
 over, the function of the mammalian heart. In practice this is performed
 by placing the pump assembly entirely within the body of the mammal and
 connecting the pump between the left ventricle and the aorta so as to
 assist left side heart function. It may also be connected to the right
 ventricle and pulmonary artery to assist the right side of the heart.
 In this instance the pump assembly includes an impeller which is fully
 sealed within the pump body and so does not require a shaft extending
 through the pump body to support it. The impeller is suspended, in use,
 within the pump body by the operation of hydrodynamic forces imparted as a
 result of the interaction between the rotating impeller, the internal pump
 walls and the fluid which the impeller causes to be urged from an inlet of
 the pump assembly to an outlet thereof.
 A preferred embodiment of the invention is the centrifugal pump 1, as
 depicted in FIGS. 1 and 2, intended for implantation into a human, in
 which case the fluid referred to below is blood. The pump housing 2, can
 be fabricated in two parts, a front part 3 in the form of a housing body
 and a back part 4 in the form of a housing cover, with a smooth join
 therebetween, for example at 5 in FIG. 1. The pump 1 has an axial inlet 6
 and a tangential outlet 7. The rotating part 100 is of very simple form,
 comprising only blades 8 and a blade support 9 to hold those blades fixed
 relative to each other. The blades may be curved as depicted in FIG. 2, or
 straight, in which case they can be either radial or back-swept, i.e. at
 an angle to the radius. This rotating part 100 will hereafter be called
 the impeller 100, but it also serves as a bearing component and as the
 rotor of a motor configuration as to be further described below whereby a
 torque is applied by electromagnetic means to the impeller 100. Note that
 the impeller has no shaft and that the fluid enters the impeller from the
 region of its axis RR. Some of the fluid passes in front of the support
 cone 9 and some behind it, so that the pump 1 can be considered of
 two-sided open type, as compared to conventional open centrifugal pumps,
 which are only open on the front side. Approximate dimensions found
 adequate for the pump 1 to perform as a ventricular assist device, when
 operating at speeds in the range 1,500 rpm to 4,000 rpm, are outer blade
 diameter 40 mm, outer housing average diameter 60 mm, and housing axial
 length 40 mm.
 As the blades 8 move within the housing, some of the fluid passes through
 the gaps, much exaggerated in FIGS. 1 and 3, between the blade edges 101
 and the housing front face 10 and housing back face 11. In all open
 centrifugal pumps, the gaps are made small because this leakage flow
 lowers the pump hydrodynamic efficiency. In the pump disclosed in this
 embodiment, the gaps are made slightly smaller than is conventional in
 order that the leakage flow can be utilised to create a hydrodynamic
 bearing. For the hydrodynamic forces to be sufficient, the blades may also
 be tapered as depicted in FIGS. 3A and 3B, so that the gap 104 is larger
 at the leading edge 102 of the blade 8 than at the trailing edge 103
 thereby providing one example of a "deformed surface" as described
 elsewhere in this specification. The fluid 105 which passes through the
 gap thus experiences a wedge shaped restriction which generates a thrust,
 as described in Reynolds' theory of lubrication (see, for example, "Modern
 Fluid Dynamics, Vol. 1 Incompressible Flow", by N. Curle and H. J. Davies,
 Van Nostrand, 1968). For blades considerably thinner than their axial
 length, the thrust is proportional to the square of the blade thickness at
 the edge, and thus in this embodiment thick blades are favoured, since if
 the proportion of the pump cavity filled by blades is constant, then the
 net thrust force will be inversely proportional to the number of blades.
 However, the blade edges can be made to extend as tails from thin blades
 as depicted in FIG. 3C in order to increase the blade area adjacent the
 walls.
 In one particular form, the tails join adjacent blades so as to form a
 complete shroud with wedges or tapers incorporated therein. An example of
 a shroud design as well as other variations on the blade structure will be
 described later in this specification.
 For manufacturing simplicity, the housing front face 10 can be made
 conical, with an angle of around 45.degree. so that it provides both axial
 and radial hydrodynamic forces. Other angles are suitable that achieve the
 functional requirements of this pump including the requirements for both
 axial and radial hydrodynamic forces.
 Other curved surfaces are possible provided both axial and radial
 hydrodynamic forces can be produced as a result of rotation of the blades
 relative to the housing surfaces.
 The housing back face 11 can include a roughly conical extension 12
 pointing into the pump cavity 106, to eliminate or minimise the effect of
 the flow stagnation point on the axis of the back housing.
 Alternatively extension 12 can resemble an impeller eye to make the flow
 mixed.
 In this preferred embodiment, for manufacturing simplicity and for
 uniformity in the flow axial direction RR, the housing back face 11 is
 made flat over the bearing surfaces, i.e. under the blade edges. With this
 the case, a slacker tolerance on the alignment between the axes of the
 front part 3 and back part 4 of the housing 2 is permissible. An
 alternative is to make the back face 11 conical at the bearing surfaces,
 with taper in the opposite direction to the front face 10, so that the
 hydrodynamic forces from the back face will also have radial components.
 Tighter tolerance on the axes alignment would then be required, and some
 of the flow would have to undergo a reversal in its axial direction. Again
 a roughly conical extension (like 12) will be needed. There may be some
 advantage in making the housing surfaces and blade edges non-straight,
 with varying tangent angle, although this will impose greater
 manufacturing complexity.
 There are several options for the shape of the taper, but in the preferred
 embodiment the amount of material removed simply varies linearly or
 approximately linearly across the blade. For the back face, the resulting
 blade edges are then planes at a slight inclination to the back face. For
 the front face, the initial blade edges are curved and the taper only
 removes a relatively small amount of material so they still appear curved.
 Alternative taper shapes can include a step in the blade edge, though the
 corner in that step would represent a stagnation line posing a thrombosis
 risk.
 For a given minimum gap, at the trailing blade edge, the hydrodynamic force
 is maximal if the gap at the leading edge is approximately double that at
 the trailing edge. Thus the taper, which equals the leading edge gap minus
 the trailing edge gap, should be chosen to match a nominal minimum gap,
 once the impeller has shifted towards that edge. Dimensions which have
 been found to give adequate thrust forces are a taper of around 0.05 mm
 for a nominal minimum gap of around 0.05 mm, and an average
 circumferential blade edge thickness of around 6 mm for 4 blades. For the
 front face, the taper is measured within the plane perpendicular to the
 axis. The axial length of the housing between the front and back faces at
 any position should then be made about 0.2 mm greater than the axial
 length of the blade, when it is coaxial with the housing, so that the
 minimum gaps are both about 0.1 mm axially when the impeller 100 is
 centrally positioned within the housing 2. Then, for example, if the
 impeller shifts axially by 0.05 mm, the minimum gaps will be 0.05 mm at
 one face and 0.15 mm at the other face. The thrust increases with
 decreasing gap and would be much larger from the 0.05 mm gap than from the
 0.15 mm gap, about 14 times larger for the above dimensions. Thus there is
 a net restoring force away from the smaller gap.
 Similarly, for radial shifts of the impeller the radial component of the
 thrust from the smaller gap on the conical housing front face would offer
 the required restoring radial force. The axial component of that force and
 its torque on the impeller would have to be balanced by an axial force and
 torque from the housing back face, and so the impeller will also have to
 shift axially and tilt its axis to be no longer parallel with the housing
 axis. Thus as the person moves and the pump is accelerated by external
 forces, the impeller will continually shift its position and alignment,
 varying the gaps in such a way that the total force and torque on the
 impeller 100 match that demanded by inertia. The gaps are so small,
 however, that the variation in hydrodynamic efficiency will be small, and
 the pumping action of the blades will be approximately the same as when
 the impeller is centrally located.
 While smaller gaps imply greater hydrodynamic efficiency and greater
 bearing thrust forces, smaller gaps also demand tighter manufacturing
 tolerances, increase frictional drag on the impeller, and impose greater
 shear stress an the fluid. Taking these points in turn, for the above 0.05
 mm tapers and gaps, tolerances of around 0.005 mm are needed, which
 imposes some cost penalty but is achievable. A tighter tolerance is
 difficult, especially if the housing is made of a plastic, given the
 changes in dimension caused by temperature and possible absorption of
 fluid by plastic materials which may be in contact with the blood such as
 Acrylic of polyurethane. The frictional drag for the above gaps produces
 much smaller torque than the typical motor torque. Finally, to estimate
 the shear stress, consider a rotation speed of 3,000 rpm and a typical
 radius of 15 mm, at which the blade speed is 4.7 ms-.sup.1 and the average
 velocity shear for an average gap of 0.075 mm is 6.2.times.10.sup.4
 s.sup.-1. For blood of dynamic viscosity 3.5.times.10.sup.-3 kgm-.sup.1
 s-.sup.1, the average shear stress would be 220 Nm.sup.-2. Other prototype
 centrifugal blood pumps with closed blades have found that slightly larger
 gaps, e.g. 0.15 mm, are acceptable for haemolysis. A major advantage of
 the open blades of the present invention is that a fluid element that does
 pass through a blade edge gap will have very short residence time in that
 gap, around 2.times.10.sup.-3 S, and the fluid element will most likely be
 swept though the pump without passing another blade edge.
 With particular reference to FIGS. 3A and 3B typical working clearances and
 working movement for the impeller 8 with respect to the upper and lower
 housing surfaces 10, 11 is of the order of 100 microns clearance at the
 top and at the bottom. In use gravitational and other forces will bias the
 impeller 8 closer to one or other of the housing walls resulting,
 typically in a clearance at one interface of the order of 50 microns and a
 corresponding larger clearance at the other interface of the order of 150
 microns. In use, likely maximum practical clearances will range from 300
 microns down to 1 micron.
 Typical restoring forces for a 25 gram rotor mass spinning at 2200 rpm are
 1.96 Newtons at a 20 micron clearance extending to 0.1 Newtons at an 80
 micron clearance.
 To minimise the net force required of the hydrodynamic bearings, the net
 axial and radial hydrodynamic forces on the impeller from the bulk fluid
 flow should be minimised, where "bulk" here means other than from the
 bearing thrust surfaces.
 The radial force on the impeller depends critically on the shape of the
 output flow collector or volute 13. The shape should be designed to
 minimise the radial impeller force over the desired range of pump speeds,
 without excessively lowering the pump efficiency. The optimal shape will
 have a roughly helical perimeter between the "cutwater" and outlet. The
 radial force can also be reduced by the introduction of an internal
 division in the volute 13 to create a second output flow collector
 passage, with tongue approximately diametrically opposite to the tongue of
 the first passage.
 An indicative plan view of impeller 100 relative to housing 2 is shown in
 FIG. 2 having a concentric volute 13.
 FIG. 17 illustrates the alternative volute arrangement comprising a split
 volute created by volute barrier 107 which causes volute 108 in a first
 hemisphere of the housing 2 to split into first half volute 109 and second
 half volute 110 over the second hemisphere. The hemispheres are defined
 respectively on each side of a diameter of the housing 2 which passes
 through or near exit point 111 of outlet 7.
 In alternative forms concentric volutes can be utilised, particularly where
 specific speed is relatively low.
 In a further particular form a vaneless diffuser may also reduce the radial
 force.
 In regard to the bulk hydrodynamic axial force, if the blade cross-section
 is made uniform in the axial direction along the rotational axis, apart
 from the conical front edge, then the pressure acting on the blade surface
 (excluding the bearing edges) will have no axial component. This also
 simplifies the blade manufacture. The blade support cone 9 must then be
 shaped to minimise axial thrust on the impeller and minimise disturbance
 to the flow over the range of speeds, while maintaining sufficient
 strength to prevent relative blade movement. The key design parameter
 affecting the axial force is the angle of the cone. The cone is drawn in
 FIG. 1 as having the same internal diameter as the blades, which may aid
 manufacture. However, the cone could be made with larger or smaller
 internal diameter to the blades. There may be advantage in using a
 non-axisymmetric support "cone", e.g. with larger radius on the trailing
 surface of a blade than the radius at the leading surface of the next
 blade. If the blades are made with non-uniform cross-section to increase
 hydrodynamic efficiency, then any bulk hydrodynamic axial force on them
 can be balanced by shaping the support cone to produce an opposite bulk
 hydrodynamic axial force on it.
 Careful design of the entire pump, employing computational fluid dynamics,
 is necessary to determine the optimal shapes of the blades 8, the volute
 13, the support cone 9 and the housing 2, in order to maximise
 hydrodynamic efficiency while keeping the bulk fluid hydrodynamic forces,
 shear and residence times low. All edges and the joins between the blades
 and the support cone should be smoothed.
 The means of providing the driving torque on the impeller 100 of the
 preferred embodiment of the invention is to encapsulate permanent magnets
 14 in the blades 8 of the impeller 100 and to drive them with a rotating
 magnetic field pattern from oscillating currents in windings 15 and 16,
 fixed relative to the housing 2. Magnets of high remanence such as
 sintered rare-earth magnets should be used to maximise motor efficiency.
 The magnets can be aligned axially but greater motor efficiency is
 achieved by tilting the magnetisation direction to an angle of around
 15.degree. to 30.degree. outwards from the inlet axis, with 22.5.degree.
 tilt suitable for a body of conical angle 45.degree.. The magnetisation
 direction must alternate in polarity for adjacent blades. Thus there must
 be an even number of blades. Since low blade number is preferred for the
 bearing force, and since two blades would not have sufficient bearing
 stiffness to rotation about an axis through the blades and perpendicular
 to the pump housing (unless the blades are very curved), four blades are
 recommended. A higher number of blades, for example 6 or 8 will also work.
 Some possible options for locating the magnets 14 within the blades 8 are
 shown in FIG. 4. The most preferred which is depicted in FIG. 4A, is for
 the blade to be made of magnet material apart from a biocompatible shell
 or coating to prevent fluid corroding the magnets and to prevent magnet
 material (which may be toxic) entering the blood stream. The coating
 should also be sufficiently durable especially at blade corners to
 withstand rubbing during start-up or during inadvertent bearing touch
 down.
 In one particular form the inside walls of the pump housing 2 are also
 coated with a biologically compatible and wear resistant material such as
 diamond coating or titanium nitride so that wear on both of the touching
 surfaces is minimised.
 An acceptable coating thickness is approximately 1 micron.
 A suitable impeller manufacturing method is to die-press the entire
 impeller, blades and support cone, as a single axially aligned magnet. The
 die-pressing is much simplified if near axially uniform blades are used
 (blades with an overhang such as in FIG. 3C are precluded). During
 pressing, the crushed rare-earth particles must be aligned in an axial
 magnetic field. This method of die-pressing with parallel alignment
 direction is cheaper for rare-earth magnets, although it produces slightly
 lower remanence magnets. The tolerance in die-pressing is poor, and
 grinding of the tapered blade edges is required. Then the magnet impeller
 can be coated, for example by physical vapour deposition, of titanium
 nitride for example, or by chemical vapour deposition, of a thin diamond
 coating or a teflon coating.
 In an alternative form the magnet material can be potted in titanium or a
 polymeric housing which is then, in turn, coated with a biologically
 compatible and tough material such as diamond coating or titanium nitride.
 Finally, to create the alternating blade polarity the impeller must be
 placed in a special pulse magnetisation fixture, with an individual coil
 surrounding each blade. The support cone of a die-pressed magnet impeller
 acquires some magnetisation near the blades, with negligible influence.
 Alternative magnet locations are sketched in FIG. 4B and FIG. 4C in which
 quadrilateral or circular cross-section magnets 14 are inserted into the
 blades. Sealing and smoothing of the blade edges over the insertion holes
 is then required to reinstate the taper.
 All edges in the pump should be radiused and surfaces smoothed to avoid
 possible damage to formed elements of the blood.
 The windings 15 and 16 of the preferred embodiment are slotless or air-gap
 windings with the same pole number as the impeller, namely four poles in
 the preferred embodiment. A ferromagnetic iron yoke 17 of conical form for
 the front winding and an iron ferromagnetic yoke 18 of annular form for
 the back winding may be placed on the outside of the windings to increase
 the magnetic flux densities and hence increase motor efficiency. The
 winding thicknesses should be designed for maximum motor efficiency, with
 the sum of their axial thicknesses somewhat less than but comparable to
 the magnet axial length. The yokes can be made of solid ferromagnetic
 material such as iron. To reduce "iron" losses, the yokes 17 can be
 laminated, for example in layers or by helically winding thin strip, or
 can be made of iron/powder epoxy composite. The yokes should be positioned
 such that there is zero net axial magnetic force on the impeller when it
 is positioned centrally in the housing. The magnetic force is unstable and
 increases linearly with axial displacement of the impeller away from the
 central position, with the gradient being called the negative stiffness of
 the magnetic force. This unstable magnetic force must be countered by the
 hydrodynamic bearings, and so the stiffness should be made as small as
 possible. Choosing the yoke thickness such that the flux density is at the
 saturation level reduces the stiffness and gives minimum mass. An
 alternative can be to have no iron yokes, completely eliminating the
 unstable axial magnetic force, but the efficiency of such designs may be
 lower and the magnetic flux density in the immediate vicinity of the pump
 may violate safety standards and produce some tissue heating. In any case,
 the stiffness is acceptably small for slotless windings with the yokes
 present. Another alternative would be to insert the windings in slots in
 laminated iron stators which would increase motor efficiency and enable
 use of less magnet material and potentially lighter impeller blades.
 However, the unstable magnetic forces would be significant for such
 slotted motors. Also, the necessity for fat blades to generate the
 required bearing forces in this embodiment allows room for large magnets,
 and so slotless windings are chosen in the preferred embodiment.
 Instead of determining the yoke positions so that the impeller has zero
 magnetic axial force in the central position, it may be possible to
 provide a bias axial magnetic force on the impeller, which can counteract
 other forces such as any average bulk hydrodynamic axial force. In
 particular, by ensuring a net axial force into the conical body, the
 thrust bearings on the cover surface can be made superfluous. However,
 such a bias would demand greater average thrust forces, smaller gaps and
 increased blood damage, and so the recommended goal is to zero both the
 magnetic and bulk hydrodynamic axial forces on the impeller when centrally
 positioned.
 The overall design requirement for exclusive hydrodynamic suspension
 requires control of the external force balance to make the relative
 magnitude of hydrodynamic thrust sufficient to overcome the external
 forces. Typical external forces include gravitational forces and net
 magnetic forces arising as a result of the motor drive.
 There are many options for the winding topology and number of phases. FIG.
 5A depicts the preferred topology for the body winding 15, viewed from the
 inlet axis.
 The cover winding 16 looks similar but the coils need not avoid the inlet
 tube and so they appear more triangular in shape. The body winding has a
 more complex three-dimensional shape with bends at the ends of the body
 cone section. Each winding consists of three coils. Each coil is made from
 a number of turns of an insulated conductor such as copper with the number
 of turns chosen to suit the desired voltage. The coil side mid-lines span
 an angle of about 50.degree.-100.degree. at the axis when the coils are in
 position. The coils for body and cover are aligned axially and the axially
 adjacent coils are connected in either parallel or series connection to
 form one phase of the three phase winding. Parallel connection offers one
 means of redundancy in that if one coil fails, the phase can still carry
 current through the other coil. In parallel connection each of the coil
 and body winding has a neutral point connection as depicted in FIG. 5A,
 whereas in series connection, only one of the windings has a neutral
 point.
 An alternative three phase winding topology, depicted in FIG. 5B, uses four
 coils per phase for each of the body and cover windings, with each coil
 wrapping around the yoke, a topology called a "Gramm ring" winding.
 Yet another three phase winding topology, depicted in FIG. 5C, uses two
 coils per phase for each of the body and cover windings, and connects the
 coil sides by azimuthal end-windings as is standard motor winding
 practice. The coils are shown tilted to approximately follow the blade
 curvature, which can increase motor efficiency, especially for the phase
 energising strategy to be described below in which only one phase is
 energised at a time. The winding construction can be simplified by laying
 the coils around pins protruding from a temporary former, the pins shown
 as dots in 2 rings of 6 pins each in FIG. 5C. The coils are labelled
 alphabetically in the order in which they would be layed, coils a and d
 for phase A, b and e for phase B, and c and f for phase C. Instead of or
 as well as pins, the coil locations could be defined by thin fins, running
 between the pins in FIG. 5C, along the boundary between the coils. The
 coil connections depicted in FIG. 5C are those appropriate for the winding
 nearest the motor terminals for the case of series connection, with the
 optional lead from the neutral point on the other winding included.
 The winding topologies depicted in FIGS. 5B and C allow the possibility of
 higher motor efficiency but only if significantly higher coil mass is
 allowed, and since option FIG. 5A is more compact and simpler to
 manufacture, it is the preferred option. Material ribs between the coils
 of option FIG. 5A can be used to stiffen the housing.
 Multi-stranded flexible conductors within a suitable biocompatible cable
 can be used to connect the motor windings to a motor controller. The
 energisation of the three phases can be performed by a standard sensorless
 controller, in which two out of six semiconducting switches in a three
 phase bridge are turned on at any one time. Alternatively, because of the
 relatively small fraction of the impeller cross-section occupied by
 magnets, it may be slightly more efficient to only activate one of the
 three phases at a time, and to return the current by a conductor from the
 neutral point in the motor. Careful attention must be paid to ensure that
 the integrity of all conductors and connections is failsafe.
 In the preferred embodiment, the two housing components 3 and 4 are made by
 injection moulding from non-electrically conducting plastic materials such
 as Lexan polycarbonate plastic. Alternatively the housing components can
 be made from ceramics. The windings and yokes are ideally encapsulated
 within the housing during fabrication moulding. In this way, the
 separation between the winding and the magnets is minimised, increasing
 the motor efficiency, and the housing is thick, increasing its mechanical
 stiffness. Alternatively, the windings can be positioned outside the
 housing, of thickness at least around 2 mm for sufficient stiffness.
 If the housing material plastic is hygroscopic or if the windings are
 outside the housing, it may be necessary to first enclose the windings and
 yoke in a very thin impermeable shell. Ideally the shell should be
 non-conducting (such as ceramic or plastic), but titanium of around 0.1 mm
 to 0.2 mm thickness would give sufficiently low eddy losses. Encapsulation
 within such a shell would be needed to prevent winding movement.
 Alternatively, the housing components 3 and 4 may be made from a
 biocompatible metallic material of low electrical conductivity, such as
 Ti-6Al-4V. To minimise the eddy current loss, the material must be as thin
 as possible, e.g. 0.1 mm to 0.5 mm, wherever the material experiences high
 alternating magnetic flux densities, such as between the coils and the
 housing inner surfaces 10 and 11.
 The combining of the motor and bearing components into the impeller in the
 preferred embodiment provides several key advantages. The rotor
 consequently has very simple form, with the only cost of the bearing being
 tight manufacturing tolerances. The rotor mass is very low, minimising the
 bearing force needed to overcome weight. Also, with the bearings and the
 motor in the same region of the rotor, the bearings forces are smaller
 than if they had to provide a torque to support magnets at an extremity of
 the rotor.
 A disadvantage of the combination of functions in the impeller is that its
 design is a coupled problem. The optimisation should ideally link the
 fluid dynamics, magnetics and bearing thrust calculations. In reality, the
 blade thickness can be first roughly sized to give adequate motor
 efficiency and sufficient bearing forces with a safety margin.
 Fortuitously, both requirements are met for four blades of approximate
 average circumferential thickness 6 mm or more. The housing, blade, and
 support cone shapes can then be designed using computational fluid
 dynamics, maintaining the above minimum average blade thickness. Finally
 the motor stator, i.e. winding and yoke, can be optimised for maximum
 motor efficiency.
 FIG. 6 depicts an alternative embodiment of the invention as an axial pump.
 The pump housing is made of two parts, a front part 19 and a back part 20,
 joined for example at 21. The pump has an axial inlet 22 and axial outlet
 23. The impeller comprises only blades 24 mounted on a support cylinder 25
 of reducing radius at each end. An important feature of this embodiment is
 that the blade edges are tapered to generate hydrodynamic thrust forces
 which suspend the impeller. These forces could be used for radial
 suspension alone from the straight section 26 of the housing, with some
 alternative means used for axial suspension, such as stable axial magnetic
 forces or a conventional tapered-land type hydrodynamic thrust bearing.
 FIG. 6 proposes a design which uses the tapered blade edges to also
 provide an axial hydrodynamic bearing. The housing is made with a reducing
 radius at its ends to form a front face 27 and a back face 28 from which
 the axial thrusts can suspend the motor axially. Magnets are embedded in
 the blades with blades having alternating polarity and four blades being
 recommended. Iron in the outer radius of the support cylinder 25 can be
 used to increase the magnet flux density. Alternatively, the magnets could
 be housed in the support cylinder and iron could be used in the blades. A
 slotless helical winding 29 is recommended, with outward bending
 end-windings 30 at one end to enable insertion of the impeller and inward
 bending windings 31 at the other end to enable insertion of the winding
 into a cylindrical magnetic yoke 32. The winding can be encapsulated in
 the back housing part 20.
 THIRD EMBODIMENT
 With reference to FIGS. 7 to 15 inclusive there is shown a further
 preferred embodiment of the pump assembly 200.
 With particular reference initially to FIG. 7 the pump assembly 200
 comprises a housing body 201 adapted for bolted connection to a housing
 cover 202 and so as to define a centrifugal pump cavity 203 therewithin.
 The cavity 203 houses an impeller 204 adapted to receive magnets 205 within
 cavities 206 defined within blades 207. As for the first embodiment the
 blades 207 are supported from a support cone 208.
 Exterior to the cavity 203 but forming part of the pump assembly 200 there
 is located a body winding 209 symmetrically mounted around inlet 210 and
 housed between the housing body 201 and a body yoke 211.
 Also forming part of the pump assembly 200 and also mounted external to
 pump cavity 203 is cover winding 212 located within winding cavity 213
 which, in turn, is located within housing cover 202 and closed by cover
 yoke 214.
 The windings 212 and 209 are supplied from the electronic controller of
 FIG. 12 as for the first embodiment the windings are arranged to receive a
 three phase electrical supply and so as to set up a rotating magnetic
 field within cavity 203 which exerts a torque on magnets 205 within the
 impeller 204 so as to urge the impeller 204 to rotate substantially about
 central axis TT of cavity 203 and in line with the longitudinal axis of
 inlet 210. The impeller 204 is caused to rotate so as to urge fluid (in
 this case blood) around volute 215 and through outlet 216.
 The assembly is bolted together in the manner indicated by screws 217. The
 yokes 211, 214 are held in place by fasteners 218. Alternatively, press
 fitting is possible provided sufficient integrity of seal can be
 maintained.
 FIG. 8 shows the impeller 204 of this embodiment and clearly shows the
 support cone 208 from which the blades 207 extend. The axial cavity 219
 which is arranged, in use, to be aligned with the longitudinal axis of
 inlet 210 and through which blood is received for urging by blades 207 is
 clearly visible.
 The cutaway view of FIG. 9 shows the axial cavity 219 and also the magnet
 cavities 206 located within each blade 207. The preferred cone structure
 220 extending from housing cover 202 aligned with the axis of inlet 210
 and axial cavity 219 of impeller 204 is also shown.
 FIG. 10 is a side section, indicative view of the impeller 204 defining the
 orientations of central axis FF, top taper edge DD and bottom taper edge
 BB, which tapers are illustrated in FIG. 11 in side section view.
 FIG. 11A is a section of a blade 207 of impeller 204 taken through plane DD
 as defined in FIG. 10 and shows the top edge 221 to be profiled from a
 leading edge 223 to a trailing edge 224 as follows: central portion 227
 comprises an ellipse with centre on the dashed midline having a semi-major
 axis of radius 113 mm and a semi-minor axis of radius 80 mm and then
 followed by leading conical surface 225 and trailing conical surface 226
 on either side thereof as illustrated in FIG. 11A. The leading surface 225
 has radius 0.05 mm less than the trailing surface 226. This prescription
 is for a taper which can be achieved by a grinding wheel, but many
 alternative prescriptions could be devised to give a taper of similar
 utility.
 The leading edge 223 is radiused as illustrated.
 FIG. 11B illustrates in cross-section the bottom edge 222 of blade 207 cut
 along plane BB of FIG. 10.
 The bottom edge includes cap 228 utilised for sealing magnet 205 within
 cavity 206.
 In this instance substantially the entire edge comprises a straight taper
 with a radius of 0.05 mm at leading edge 229 and a radius of 0.25 mm at
 trailing edge 230.
 The blade 207 is 6.0 mm in width excluding the radii at either end.
 FIG. 12 comprises a block diagram of the electrical controller suitable for
 driving the pump assembly 200 and comprises a three phase commutation
 controller 232 adapted to drive the windings 209, 212 of the pump
 assembly. The commutation controller 232 determines relative phase and
 frequency values for driving the windings with reference to set point
 speed input 233 derived from physiological controller 234 which, in turn,
 receives control inputs 235 comprising motor current input and motor speed
 (derived from the commutation controller 232), patient blood flow 236, and
 venous oxygen saturation 237. The pump blood flow can be approximately
 inferred from the motor speed and current via curve-fitted formulae.
 FIG. 13 is a graph of pressure against flow for the pump assembly 200 where
 the fluid pumped is 18% glycerol for impeller rotation velocity over the
 range 1500 RPM to 2500 RPM. The 18% glycerol liquid is believed to be a
 good analogue for blood under certain circumstances, for example in the
 housing gap.
 FIG. 14 graphs pump efficiency against flow for the same fluid over the
 same speed ranges as for FIG. 13.
 FIG. 15 is a graph of electrical power consumption against flow for the
 same fluid over the same speed ranges as for FIG. 13.
 The common theme running through the first, second and third embodiments
 described thus far is the inclusion in the impeller of a taper or other
 deformed surface which, in use, moves relative to the adjacent housing
 wall thereby to cause a restriction with respect to the line of movement
 of the taper or deformity thereby to generate thrust upon the impeller
 which includes a component substantially normal to the line of movement of
 the surface and also normal to the adjacent internal pump wall with
 respect to which the restriction is defined for fluid located
 therebetween.
 In order to provide both radial and axial direction control at least one
 set of surfaces must be angled with respect to the longitudinal axis of
 the impeller (preferably at approximately 45.degree. thereto) thereby to
 generate or resolve opposed radial forces and an axial force which can be
 balanced by a corresponding axial force generated by at least one other
 tapered or deformed surface located elsewhere on the impeller.
 In the forms thus far described top surfaces of the blades 8, 207 are
 angled at approximately 45.degree. with respect to the longitudinal axis
 of the impeller 100, 204 and arranged for rotation with respect to the
 internal walls of a similarly angled conical pump housing. The top
 surfaces of the blades are deformed so as to create the necessary
 restriction in the gap between the top surfaces of the blades and the
 internal walls of the conical pump housing thereby to generate a thrust
 which can be resolved to both radial and axial components.
 In the examples thus far the bottom faces of the blades 8, 207 comprise
 surfaces substantially lying in a plane at right angles to the axis of
 rotation of the impeller and, with their deformities define a gap with
 respect to a lower inside face of the pump housing against which a
 substantially only axial thrust is generated.
 Other arrangements are possible which will also, relying on these
 principles, provide the necessary balanced radial and axial forces. Such
 arrangements can include a double cone arrangement where the conical top
 surface of the blades is mirrored in a corresponding bottom conical
 surface. The only concern with this arrangement is the increased depth of
 pump which can be a problem for in vivo applications where size
 minimisation is an important criteria.
 FOURTH EMBODIMENT
 With reference to FIG. 18 a further embodiment of the invention is
 illustrated comprising a plan view of the impeller 300 forming part of a
 "channel" pump. In this embodiment the blades 301 have been widened
 relative to the blades 207 of the third embodiment to the point where they
 are almost sector-shaped and the flow gaps between adjacent blades 301, as
 a result, take the form of a channel 302, all in communication with axial
 cavity 303.
 A further modification of this arrangement is illustrated in FIG. 19
 wherein impeller 304 includes sector-shaped blades 305 having curved
 leading and trailing portions 306, 307 respectively thereby defining
 channels 308 having fluted exit portions 309.
 As with the first and second embodiments the radial and axial hydrodynamic
 forces are generated by appropriate profiling of the top and bottom faces
 of the blades 301, 305 (not shown in FIGS. 18 and 19).
 FIG. 20 illustrates a perspective view of an impeller 304 which follows the
 theme of the impeller arrangement of FIGS. 18 and 19 in perspective view
 and where like parts are numbered as for FIG. 19. In this case the four
 blades 305 are joined at mid-portions thereof by a blade support in the
 form of a conical rim 350 and have edge portions which are shaped so as to
 have an increased curvature on the trailing edge 351 thereof compared with
 the leading edge 352.
 FIFTH EMBODIMENT
 A fifth embodiment of a pump assembly according to the invention comprises
 an impeller 410 as illustrated in FIG. 21 where, conceptually, the upper
 and lower surfaces of the blades of previous embodiments are
 interconnected by a top shroud 411 and a bottom shroud 412. In this
 embodiment the blades 413 can be reduced to a very small width as the
 hydrodynamic behaviour imparted by their surfaces in previous embodiments
 is now given effect by the profiling of the shrouds 411, 412 which, in
 this instance, comprises a series of smooth-edged wedges with the leading
 surface of one wedge directly interconnected to the trailing edge of the
 next leading wedge 414.
 As for previous embodiments the top shroud 411 is of overall conical shape
 thereby to impart both radial and axial thrust forces whilst the bottom
 shroud 412 is substantially planar thereby to impart substantially only
 axial thrust forces.
 It is to be understood that, whilst the example of FIG. 21 shows the
 surfaces of the shroud 411 angled at approximately 45.degree. to the
 vertical, other inclinations are possible extending to an inclination of
 0.degree. to the vertical which is to say the impeller 410 can take the
 form of a cylinder with surface rippling or other deformations which
 impart the necessary hydrodynamic lift, in use.
 With reference to FIGS. 22 to 24 a specific example of the concept embodied
 in FIG. 21 is illustrated and wherein like components are numbered as for
 FIG. 21.
 It will be observed that, with reference to FIG. 24, the blades 413 are
 thin compared to previous embodiments and, in this instance, are arcuate
 channels 416 therebetween which allow fluid communication from a centre
 volume 417 to the periphery 418 of the impeller 410.
 In this arrangement it will be noted that the wedges 414 are separated one
 from the other on each shroud by channels 419. The channels extend
 radially down the shroud from the centre volume 417 to the periphery 418.
 In such designs with thin blades, the magnets required for the driving
 torque can be contained within the top or bottom volute or both, along
 with the optional soft magnetic yokes to increase motor efficiency.
 A variation of this embodiment is to have the wedge profiling cut into the
 inner surfaces of the housing and have smooth shroud surfaces.
 SIXTH EMBODIMENT
 In contrast to the embodiments illustrated with respect to FIGS. 3A, 3B and
 3C an arrangement is shown in FIG. 25 wherein the "deformed surface"
 comprises a stepped formation 510 forming part of an inner wall of the
 pump housing (not shown). In this instance the rotor including blade 511
 includes a flat working surface 512 (and not having a deformed surface
 therein) which is adapted for relative movement in the direction of the
 arrow shown with respect to the stepped formation 510 thereby to generate
 hydrodynamic thrust therebetween.
 SEVENTH EMBODIMENT
 With reference to FIG. 26 there is shown an arrangement of rotor blade 610
 with respect to stepped formation 611 and wherein the rotor blade 610
 includes a deformed surface 612 at a working face thereof. In this
 instance the deformation comprises curved edges 613, 614. As for the
 previous embodiment relative movement of the rotor blade 610 in the
 direction of the arrow with respect to deformed surface 611 forming part
 of the pump housing (not shown) causes relative hydrodynamic thrust
 therebetween.
 The foregoing describes principles and examples of the present invention,
 and modifications, obvious to those skilled in the art, can be made
 thereto without departing from the scope and spirit of the invention.
 PRINCIPLES OF OPERATION
 With particular reference to FIG. 27 this specification describes the
 suspension of an impeller 600 within a pump housing 601 by the use of
 hydrodynamic forces. In this specification the suspension of the impeller
 600 is performed dominantly which is to say exclusively by hydrodynamic
 forces.
 The hydrodynamic forces are forces which are created by relative movement
 between two surfaces which have a fluid in the gap between the two
 surfaces. In the case of the use of the pump assembly 602 as a rotary
 blood pump the fluid is blood.
 The hydrodynamic forces can arise during relative movement between two
 surfaces even where those surfaces are substantially entirely parallel to
 each other or non-deformed. However, in this specification, hydrodynamic
 forces are caused to arise during relative movement between two surfaces
 where at least one of the surfaces includes a "deformed surface".
 In this specification "deformed surface" means a surface which includes an
 irregularity relative to a surface which it faces such that, when the
 surface moves in a predetermined direction relative to the surface which
 it faces the fluid located in the gap there between experiences a change
 in relative distance between the surfaces along the line of movement
 thereby to cause a hydrodynamic force to arise therebetween in the form of
 a thrust force including at least a component substantially normal to the
 plane of the gap defined at any given point between the facing surfaces.
 In the example of FIG. 27 there is a first deformed surface 603 forming at
 least part of a first face 604 of impeller 600 and a second deformed
 surface 605 on a second face 606 of the impeller 600.
 The inset of FIG. 27 illustrates conceptually how the first deformed
 surface 603 may form only part of the first face 604.
 The first deformed surface 603 faces first inner surface 607 of the pump
 housing 601 whilst second deformed surface 605 faces second inner surface
 608 of the pump housing 601.
 In use first gap 609 defined between first deformed surface 603 and first
 inner surface 607 has a fluid comprising blood located therein whilst
 second gap 610 defined between second deformed surface 605 and second
 inner surface 608 also has a fluid comprising blood located therein.
 In use impeller 600 is caused to rotate about impeller axis 611 such that
 relative movement across first gap 609 between first deformed surface 603
 and first inner face 607 occurs and also relative movement across second
 gap 610 between second deformed surface 605 and second inner surface 608
 occurs. The orientation of the deformities of first deformed surface 603
 and second deformed surface 605 relative to the line of movement of the
 deformed surfaces 603, 605 relative to the inner surfaces 607, 608 is such
 that the fluid in the gaps 609, 610 experiences a change in height of the
 gap 609, 610 as a function of time and with the rate of change dependant
 on the shape of the deformities of the deformed surfaces and also the rate
 of rotation of the impeller 600 relative to the housing 601. That is, at
 any given point on either inner surface 607 or 608, the height of the gap
 between the inner surface 607 or 608 and corresponding deformed surface
 603 or 605 will vary with time due to passage of the deformed surface 603
 or 605 over the inner surface.
 Hydrodynamic forces in the form of thrust forces normal to the line of
 relative movement of the respective deformed surfaces 603, 605 relative to
 the inner surfaces 607, 608 thus arise.
 With this configuration it will be noted that the first gap 609 lies
 substantially in a single plane whilst the second gap 610 is in the form
 of a cone and angled at an acute angle relative to the plane of the first
 gap 609.
 Accordingly, the thrust forces which can be enlisted to first gap 609 and
 second gap 610 are substantially normal to and distributed across both the
 predominantly flat plane of first deformed surface 603 and normal to the
 substantially conical surface of second deformed surface 605 thereby
 permitting restoring forces to be applied between the impeller 600 and the
 pump housing 601 thereby to resist forces which seek to translate the
 impeller 600 in space relative to the pump housing 601 and also to rotate
 the impeller 600 about any axis (other than about the impeller axis 611)
 relative to the pump housing 601. This arrangement substantially resists
 five degrees of freedom of movement of impeller 600 with respect to the
 housing 601 and does so predominantly without any external intervention to
 control the position of the impeller with respect to the housing given
 that disturbing forces from other sources, most notably magnetic forces on
 the impeller due to its use as rotor of the motor are net zero when the
 impeller occupies a suitable equilibrium position. The balance of all
 forces on the rotor, effected by manipulation of magnetic and other
 external sources, may be adjusted such that the rotor is predominantly
 hydrodynamically born.
 It will be observed that these forces increase as the gaps 609, 610 narrow
 relative to a defined operating position and decrease as the gaps 609, 610
 increase relative to a defined operating gap. Because of the opposed
 orientation of first deformed surface 603 relative to second deformed
 surface 605 it is possible to design for an equilibrium position of the
 impeller 600 within the pump housing 601 at a defined equilibrium gap
 distance for gaps 609, 610 at a specified rotor rotational speed about
 axis 611 and rotor mass leading to a close approximation to an
 unconditionally stable environment for the impeller 600 within the pump
 housing 601 against a range of disturbing forces.
 Characteristics and advantages which flow from the arrangement described
 above and with reference to the embodiments includes:
 1. Low haemolysis, hence low running speed and controlled fluid dynamics
 (especially shear stress) in the gap between the casing and impeller. This
 in turn led to the selection of radial off-flow and minimal incidence at
 on-flow to the rotor;
 2. Radial or near-radial off-flow from the impeller can be chosen in order
 to yield a "flat" pump characteristic (HQ) curve.
 INDUSTRIAL APPLICABILITY
 The pump assembly 1, 200 is applicable to pump fluids such as blood on a
 continuous basis. With its expected reliability it is particularly
 applicable as an in vivo heart assist pump.
 The pump assembly can also be used with advantage for the pumping of other
 fluids where damage to the fluid due to high shear stresses must be
 avoided or where leakage of the fluid must be prevented with a very high
 degree of reliability--for example where the fluid is a dangerous fluid.