U.S. patent application number 10/921662 was filed with the patent office on 2005-02-03 for rotary blood pump and control system therefor.
Invention is credited to Ayre, Peter Joseph, Tansley, Geoffrey Douglas, Watterson, Peter Andrew, Woodard, John Campbell.
Application Number | 20050025630 10/921662 |
Document ID | / |
Family ID | 3814158 |
Filed Date | 2005-02-03 |
United States Patent
Application |
20050025630 |
Kind Code |
A1 |
Ayre, Peter Joseph ; et
al. |
February 3, 2005 |
Rotary blood pump and control system therefor
Abstract
A pump assembly and estimation and control system therefor, the
pump adapted for continuous flow pumping of blood. In a particular
form, the pump is a centrifugal pump wherein the impeller is
entirely sealed within the pump housing and is exclusively
hydrodynamically suspended therein against movement in three
translational and two rotational degrees of freedom as the impeller
rotates within the fluid urged by electromagnetic means external to
the pump cavity. Hydrodynamic suspension is assisted by the
impeller having deformities therein such as blades with surfaces
tapered from the leading edges to the trailing edges of bottom and
top surfaces thereof.
Inventors: |
Ayre, Peter Joseph; (Crows
Nest, AU) ; Tansley, Geoffrey Douglas; (Mi-Colah,
AU) ; Watterson, Peter Andrew; (West Ryde, AU)
; Woodard, John Campbell; (Thornleigh, AU) |
Correspondence
Address: |
KNOBBE MARTENS OLSON & BEAR LLP
2040 MAIN STREET
FOURTEENTH FLOOR
IRVINE
CA
92614
US
|
Family ID: |
3814158 |
Appl. No.: |
10/921662 |
Filed: |
August 19, 2004 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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10921662 |
Aug 19, 2004 |
|
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09980682 |
Aug 15, 2002 |
|
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09980682 |
Aug 15, 2002 |
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PCT/AU00/00355 |
Apr 20, 2000 |
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Current U.S.
Class: |
417/53 ;
417/423.12 |
Current CPC
Class: |
A61M 60/422 20210101;
A61M 60/824 20210101; A61M 60/205 20210101; Y10S 415/90 20130101;
A61M 60/82 20210101; A61M 60/50 20210101; A61M 60/148 20210101;
A61M 2205/3334 20130101 |
Class at
Publication: |
417/053 ;
417/423.12 |
International
Class: |
F04B 001/00 |
Foreign Application Data
Date |
Code |
Application Number |
Apr 23, 1999 |
AU |
PP 9959 |
Claims
What is claimed is:
1. A method of hydrodynamically suspending and controlling an
impeller within a rotary pump for support in at least one of a
radial and axial direction; the method comprising: incorporating a
deformed surface in at least part of the impeller so that, in use,
a thrust is created between the deformed surface and an adjacent
pump housing during relative movement therebetween; and maintaining
the speed of rotation of the impeller within a range whereby the
impeller, in use, substantially resists five degrees of freedom of
movement with respect to the pump housing without any external
intervention.
2. The method of claim 1 wherein the deformed surface includes a
taper.
3. The method of claim 2 wherein the taper is arranged so that
there is a larger gap at a leading edge thereof between the
impeller and the pump housing than at a trailing edge thereof.
4. An estimation and control system for a pump; the pump of the
type having an impeller located within a pump cavity in a pump
housing; the housing having a fluid inlet in fluid communication
with the cavity; the housing having a fluid outlet in fluid
communication with the pump cavity; the impeller urged to rotate
about an impeller axis so as to cause fluid to be urged from the
inlet through the pump cavity to the pump outlet; the impeller
urged to rotate by an impeller drive; the impeller supported for
rotational movement by an impeller support; the pump maintained at
or near a predetermined operating point by a controller acting on
the impeller drive; the controller receiving as input variables at
least a first input variable derived from the impeller drive; the
controller receiving at least a second input variable also derived
from the impeller drive; the controller thereby calculating an
estimate of the operating point to an approximation of
predetermined accuracy relying on signals available from the
impeller drive; the controller controlling the pump by comparing
the predetermined operating point with the estimate of the
operating point; and wherein the pump is arranged to operate
according to a relatively flat pressure versus flow rate
characteristic.
5. The estimation and control system of claim 4 wherein there is no
inflexion point in the pressure versus flow rate characteristic at
or near the predetermined operating point.
6. The estimation and control system of claim 4 wherein the pump
includes near-radial off-flow from the impeller.
7. The estimation and control system of claim 4 wherein the pump
has a low specific speed.
Description
RELATED APPLICATIONS
[0001] This application is a divisional application of U.S.
application Ser. No. 09/980,682 filed Aug. 15, 2002 which is a
national phase application of the International Application
PCT/AU00/00355 filed Apr. 20, 2000 and claims the priority benefits
of U.S. application Ser. No. 09/299,038 filed Apr. 23, 1999 (issued
Jun. 26, 2001 as U.S. Pat. No. 6,250,880) and Australian PP 9959
filed Apr. 23, 1999, all of which are incorporated herein by
reference in their entireties.
BACKGROUND OF THE INVENTION
[0002] 1. Field of the Invention
[0003] 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 fixed relative to the current pattern
generated in coils pump housing.
[0004] In addition, a non-contact estimation and control system is
described for use in conjunction with the rotary pumps of the
invention.
[0005] 2. Description of the Related Art
[0006] 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.
[0007] 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.
[0008] 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.
[0009] 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 -3 proteins, with the risk of
embolisation and bearing seizure.
[0010] 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 Eamshaw'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.
[0011] 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.
[0012] 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.
[0013] U.S. Pat. No. 4,944,748 discloses a magnetically suspended
impeller within a pump. It does not disclose an exclusively
hydrodynamically suspended impeller within a pump.
[0014] U.S. Pat. No. 4,688,998 again discloses a magnetically
suspended impeller. It does not disclose a hydrodynamically
suspended impeller, much less an exclusively hydrodynamically
suspended impeller within a pump.
[0015] WO 91/19103 to NU-TECH discloses an axial flow blood pump
having a hydrodynamically suspended rotor assisted by magnetic or
mechanical stabilisation.
[0016] U.S. Pat. No. 5,112,200 to NU-TECH discloses hydrodynamic
support in at least one dimension, but utilising prior art
hydrodynamic lift surfaces which do not include the deformed
surfaces of the present invention.
[0017] WO 94/13955 discloses a fluid pump which relies on a
magnetically levitated impeller.
[0018] U.S. Pat. No. 4,382,199 to NU-TECH discloses a rotor and
impeller combination which employs "squeeze film effects, dash pot
effects and hydrodynamic effects, all of which combine and
co-operate to prevent metal-to-metal contact between the rotor and
the stator and to lubricate the rotor as it rotates within the
stator" (column 6). There is no disclosure of exclusive
hydrodynamic support in all dimensions by the use of deformed
surfaces.
[0019] 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.
[0020] 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.
[0021] 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
[0022] Accordingly, in one broad form of the invention there is
provided 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; and wherein at least one of said impeller of said
housing includes at least a first deformed surface lying on at
least part of a first face and a second deformed surface lying on
at least part of a second face which, in use, move relative to
respective facing surfaces on the other of said impeller or said
housing thereby to form at least two relatively moving surfaces
pairs which 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 first deformed surface and said second deformed
surface with respect to said facing surfaces; and wherein 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.
[0023] In yet a further broad form of the invention there is
provided an estimation and control system for a pump; said pump of
the type having an impeller located within a pump cavity in a pump
housing; said housing having a fluid inlet in fluid communication
with said pump cavity; said impeller urged to rotate about an
impeller axis so as to cause fluid to be urged from said inlet
through said pump cavity to said pump outlet; said impeller urged
to rotate by impeller urging means; said impeller supported for
rotational movement by impeller support means; said impeller
maintained at or near a predetermined speed of rotation by control
means acting on said impeller urging means; said control means
receiving as input variables a first input variable comprising
power consumed by said urging means; said control means receiving a
second input variable comprising actual speed of rotation of said
impeller; said control means thereby estimating head across the
pump and/or rate of flow of said fluid to an approximation of
predetermined accuracy relying on signals available from said
urging means; said control system adapted to maintain speed of
rotation of said impeller within a range whereby said impeller, in
use, substantially resists five degrees of freedom of movement with
respect to said pump housing predominantly without any external
intervention from said control system to control the position of
said impeller with respect to said housing.
[0024] In yet a further broad form of the invention there is
provided a rotary blood pump and an estimation and control system
therefor, said pump having an impeller suspended hydrodynamically
within a pump housing by thrust forces generated by the impeller
during movement in use of the impeller as it rotates about an
impeller axis; said estimation and control system of the type
described above.
[0025] In yet 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 impeller axis to cause a
pressure differential between an inlet side of the pump housing of
said pump and an outlet side of the pump housing of said pump; said
impeller suspended hydrodynamically by thrust forces generated by
the impeller during movement in use of the impeller; said pump
controlled by the estimation and control system as described
above.
[0026] In yet a further broad form of the invention there is
provided a seal-less, shaft-less pump comprising a housing defining
a chamber therein and having a liquid inlet to said chamber and a
liquid outlet from said chamber; said pump further including an
impeller located within said chamber; the arrangement between said
impeller, said inlet, said outlet, and the internal walls of said
chamber being such that upon rotation of said impeller about an
impeller axis relative to said housing liquid is urged from said
inlet through said chamber to said outlet; and wherein thrust
forces are generated by relative movement of said impeller with
respect to said housing; said pump controlled by the estimation and
control system as described above.
[0027] In yet a further broad form of the invention there is
provided a 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 in at least one of a radial or axial direction by
thrust forces generated by the impeller during movement in use of
the impeller; said pump controlled by the estimation and control
system as described above.
[0028] In yet a further broad form of the invention there is
provided a method of hydrodynamically suspending and controlling an
impeller within a rotary pump for support in at lest one of a
radial or axial direction; said method comprising incorporating a
deformed surface in at least part of said impeller so that, in use,
a thrust is created between said deformed surface and the adjacent
pump casing during relative movement therebetween; said method
further including the step of maintaining speed of rotation of said
impeller within a range whereby said impeller, in use,
substantially resists five degrees of freedom of movement with
respect to said pump housing without any external intervention.
[0029] In yet a further broad form of the invention there is
provided an estimation and control system for a pump; said pump of
the type having an impeller located within a pump cavity in a pump
housing; said housing having a fluid inlet in fluid communication
with said cavity; said housing having a fluid outlet in fluid
communication with said pump cavity; said impeller urged to rotate
about an impeller axis so as to cause fluid to be urged from said
inlet through said pump cavity to said pump outlet; said impeller
urged to rotate by impeller urging means; said impeller supported
for rotational movement by impeller support means; said pump
maintained at or near a predetermined operating point by control
means acting on said impeller urging means; said control means
receiving as input at least a first input variable derived from
said urging means; said control means receiving at least a second
input variable also derived from said urging means; said control
means thereby calculating an estimate of said operating point to an
approximation of predetermined accuracy relying on signals
available from said urging means; said control means controlling
said pump by comparing said predetermined operating point with said
estimate of said operating point; and wherein instantaneous pump
speed and electrical input power are allowed to be modulated by the
heart, in use, by appropriate selection of a control time
constant.
[0030] In yet a further broad form of the invention there is
provided a physiological controller for use in association with a
pump; said controller monitoring estimated flow of fluid within
said pump and pressure across said pump by non-contact means
thereby to control speed of rotation of an impeller within said
pump; and wherein said controller permits impeller speed to vary
under a pulsating fluid load thereby to assist in calculation and
adjustment of impeller speed set point.
[0031] Preferably said pump comprises a ventricular assist device
adapted to assist operation of a ventricle of a hear and wherein
said control means adjusts pump output so that, in alternation
fashion, said ventricle in conjunction with said aortic valve is
allowed to eject blood over a predetermined number of cardiac
cycles and then said ventricle in conjunction with said aortic
valve is caused to eject blood over a following predetermined
number of cardiac cycles.
[0032] In yet a further broad form of the invention there is
provided an estimation and control system for a pump; said pump of
the type having an impeller located within a pump cavity in a pump
housing; said housing having a fluid inlet in fluid communication
with said cavity; said housing having a fluid outlet in fluid
communication with said pump cavity; said impeller urged to rotate
about an impeller axis so as to cause fluid to be urged from said
inlet through said pump cavity to said pump outlet; said impeller
urged to rotate by impeller urging means; said impeller supported
for rotational movement by impeller support means; said pump
maintained at or near a predetermined operating point by control
means acting on said impeller urging means; said control means
receiving as input variables at least a first input variable
derived from said urging means; said control means receiving at
least a second input variable also derived from said urging means;
said control means thereby calculating an estimate of said
operating point to an approximation of predetermined accuracy
relying on signals available from said urging means; said control
means controlling said pump by comparing said predetermined
operating point with said estimate of said operating point; and
wherein said pump is arranged to operate according to a relatively
flat HQ characteristic.
[0033] Preferable there is no inflexion point of said HQ
characteristic at or near said predetermined operating point.
[0034] Preferably said pump includes near-radial off-flow from said
impeller.
[0035] Preferably said pump has a low specific speed.
[0036] Preferably said pump is a low specific speed pump.
[0037] Preferably said pump is specified in a range of 100-2000
rev/min (gal/min).sup.1/2ft.sup.-3/4.
[0038] Preferably said pump has a specific speed of approximately
900-1000 rev/min (gal/min).sup.1/2ft.sup.-3/4.
[0039] Preferably instantaneous pump speed and electrical input
power are allowed to be modulated by the heart, in use, by
appropriate selection of time constant.
[0040] Preferably the time constant of the control system is
greater than the rotational, inertial time constant of the
impeller.
[0041] Preferably said time constant is at least one cardiac
cycle.
[0042] Preferably said first input variable comprises instantaneous
pump speed.
[0043] Preferably said second input variable comprises electrical
input power to said impeller urging means.
[0044] Preferably said pump is arranged to operate according to a
relatively flat HQ characteristic.
[0045] In a particular preferred form said HQ characteristic is
sufficiently flat that head will remain constant to a sufficient
approximation over a predetermined operating range whereby, over
said operating range whereby, over said operating range, said
system can assume that pump speed will be proportional to flow
rate.
[0046] Preferably said predetermined operating point is calculated
so as to maintain minimum pump speed such that the minimum head
pressure across the pump does not increase.
[0047] Preferably said system ensures that minimum pump speed is
always greater than or equal to the minimum speed at which
non-regurgitant flow will occur.
[0048] Preferably the speed at which regurgitant or negative flow
will begin to occur is determined as that pump set point speed
where levels and phase lags between pump outlet and inlet pressures
fall during diastole cause flow reversal.
[0049] In a particular preferred form the pump speed at which
regurgitation is calculated to occur is calculated according
to:
Nregurg=N(t) for Qdiastole=0L/min
BRIEF DESCRIPTION OF THE DRAWINGS
[0050] Embodiments of the present invention will now be described,
with reference to the accompanying drawings, wherein.
[0051] FIG. 1 is a longitudinal cross-sectional view of a preferred
embodiment of the invention;
[0052] FIG. 2 is a cross-sectional view taken generally along the
line Z-Z of FIG. 1;
[0053] FIG. 3A is a cross-sectional view of an impeller blade taken
generally along the line A-A of FIG. 2;
[0054] FIG. 3B is an enlargement of the blade-pump housing
interface portion of FIG. 3A;
[0055] FIG. 3C is an alternative impeller blade shape;
[0056] FIGS. 4A, B, C illustrate various possible locations of
magnet material within a blade;
[0057] FIGS. 5A, B and C are left-hand end views of possible
winding geometries taken generally along the line S-S of FIG.
1;
[0058] FIG. 6 is a diagrammatic cross-sectional view of an
alternative embodiment of the invention as an axial pump;
[0059] FIG. 7 is an exploded, perspective view of a centrifugal
pump assembly according to a further embodiment of the
invention;
[0060] FIG. 8 is a perspective view of the impeller of the assembly
of FIG. 7;
[0061] FIG. 9 is a perspective, cut away view of the impeller of
FIG. 8 within the pump assembly of FIG. 7;
[0062] FIG. 10 is a side section indicative view of the impeller of
FIG. 8;
[0063] FIG. 11 is a detailed view in side section of blade portions
of the impeller of FIG. 10;
[0064] FIG. 12 is a block diagram of an electronic driver circuit
for the pump assembly of FIG. 7;
[0065] FIG. 13 is a graph of head versus flow for the pump assembly
of FIG. 7;
[0066] FIG. 14 is a graph of pump efficiency versus flow for the
pump assembly of FIG. 7;
[0067] FIG. 15 is a graph of electrical power consumption versus
flow for the pump assembly of FIG. 7;
[0068] FIG. 16 is a plan, section view of the pump assembly showing
a volute arrangement according to a preferred embodiment;
[0069] FIG. 17 is a plan, section view of a pump assembly showing
an alternative volute arrangement;
[0070] FIG. 18 is a plan view of an impeller according to a further
embodiment of the invention;
[0071] FIG. 19 is a plan view of an impeller according to a further
embodiment of the invention;
[0072] FIG. 20 is a perspective view of an impeller according to a
further embodiment of the invention;
[0073] FIG. 21 is a perspective view of an impeller according to
yet a further embodiment of the invention;
[0074] FIG. 22 is a perspective, partially cut away view of an
impeller according to yet a further embodiment of the
invention;
[0075] FIG. 23 is a top, perspective view of the impeller of FIG.
22;
[0076] FIG. 24 is a perspective view of the impeller of FIG. 22
with its top shroud removed;
[0077] FIG. 25 illustrates an alternative embodiment wherein the
deformed surface is located on the pump housing; and
[0078] FIG. 26 illustrates a further embodiment wherein deformed
surfaces are located both on the impeller and on the housing.
[0079] FIG. 27 illustrates diagrammatically the basis of operation
of the "deformed surfaces" utilised for hydrodynamic suspension of
embodiments of the invention.
[0080] FIG. 28 is a block diagram of a non-contact estimation and
control system in accordance with a first embodiment of the
invention applied to a blood pump;
[0081] FIG. 29 is a characteristic estimation curve utilised by the
non-contact estimation and control system of FIG. 1;
[0082] FIG. 30 is a side, cut-away view of the pump of FIG. 1;
[0083] FIG. 31 is a plan, cut-away view of the coil and magnet
system of the pump of FIG. 1;
[0084] FIG. 32 is a block diagram of an electronic driver circuit
for the pump assembly of FIG. 7;
[0085] FIG. 33 illustrates efficiency versus specific speed for a
range of pump types, to be contrasted with the flat HQ curves of
FIG. 13;
[0086] FIG. 34 provides a graphical comparison of HQ curves for
pump constructions according to embodiments of the invention
compared with typical centrifugal pump HQ curves;
[0087] FIG. 35 illustrates a particular preferred form of impeller
described with reference to example 2;
[0088] FIG. 36 illustrates an implanted rotary pump assembly and
associated control system according to example 2;
[0089] FIG. 37 illustrates graphically a control strategy to avoid
over pumping for the system of example 2;
[0090] FIG. 38 is a graphical illustration of application of the
algorithms of the control system to estimate pressure head for
example 2; and
[0091] FIG. 39 is a graphical illustration of application of the
algorithms of the control system to provide a flow rate estimate
for the system of example 2.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT
[0092] 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.
[0093] 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.
[0094] 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 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.
[0095] 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 bearing faces 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 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 wedge-shaped
restriction defined by at least one "deformed surface" as described
elsewhere in this specification and a corresponding opposing
surface. 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 length, the thrust is proportional to the square
of the blade thickness at the bearing face, 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 bearing faces can be made to extend as tails
from thin blades as depicted in FIG. 3C in order to increase the
blade face area adjacent the walls.
[0096] 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.
[0097] For manufacturing simplicity, the housing front face 10 can
be made conical, with an angle of around 450 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.
[0098] 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.
[0099] In one form the housing back face 11 may 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.
[0100] Alternatively extension 12 can resemble an impeller eye to
make the flow mixed.
[0101] In an alternative form the extension 12 can be omitted for
ease of manufacture.
[0102] 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 bearing faces. 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.
[0103] There are many profiles of bearing surface which will
generate the wedge-shaped restriction. In the preferred embodiment
the amount of material removed simply varies linearly or
approximately linearly across the blade between the body and
trailing edges. Alternative taper shapes can include a radiused
leading edge or a step in the blade bearing face, though the corner
in that step may represent a stagnation line posing a thrombosis
risk.
[0104] For a given minimum gap, at the trailing blade edge, the
hydrodynamic force is maximal if the gap at the leading edge of the
blade end face is approximately double that at the trailing edge of
the blade end face. Thus the taper, which equals the blade face
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 bearing face
thickness of around 60 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.
[0105] 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.
[0106] 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.1s-.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 bearing face 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 bearing face.
[0107] 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.
[0108] 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.
[0109] 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.
[0110] 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.
[0111] An indicative plan view of impeller 100 relative to housing
2 is shown in FIG. 2 having a concentric volute 13.
[0112] 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.
[0113] In alternative forms concentric volutes can be utilised,
particularly where specific speed is relatively low.
[0114] In a further particular form a vaneless diffuser may also
reduce the radial force.
[0115] 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 surface, then
the pressure acting on the blade surface (excluding the bearing
surfaces) will have no axial component. This also simplifies the
blade manufacture. The blade support 9 should 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 support.
The support is drawn in FIG. 1 as having the same internal diameter
as the blades, which may aid manufacture. However, the support
could be made with larger or smaller internal diameter to the
blades. There may be advantage in using a non-axisymmetric support,
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 to produce an opposite bulk
hydrodynamic axial force on it.
[0116] Alternatively, by careful manufacture of taper axial thrust
can be engineered.
[0117] 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 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 should be
smoothed.
[0118] 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.
[0119] 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.
[0120] 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 titanium nitride so that wear on both of the
touching surfaces is minimised.
[0121] An acceptable coating thickness is approximately 1
micron.
[0122] In one 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 titanium
nitride.
[0123] In an alternative form a suitable impeller manufacturing
method is to die-press the entire impeller, blades and support, 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 surfaces 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 teflon coating.
[0124] Finally, to create the alternating blade polarity the
impeller may be placed in a special pulse magnetisation fixture,
with an individual coil surrounding each blade. The support of a
die-pressed magnet impeller acquires some magnetisation near the
blades, with negligible influence.
[0125] 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
bearing surfaces over the insertion holes is then required to
reinstate the taper.
[0126] All edges in the pump should be radiused and surfaces
smoothed to avoid possible damage to formed elements of the
blood.
[0127] 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.
[0128] 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.
[0129] 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.
[0130] 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.
[0131] 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 support 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.
[0132] An alternative three phase winding topology, depicted in
FIG. 5S, 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.
[0133] 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 labeled
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.
[0134] 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.
[0135] 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.
[0136] In one 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.
[0137] 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). Titanium of around 0.1 mm to 0.2 mm thickness gives
sufficiently low eddy losses. Encapsulation within such a shell is
needed to prevent winding movement.
[0138] Alternatively, and in a particularly preferred embodiment
the housing components 3 and 4 may be made from a biocompatible
metallic material of low electrical conductivity, such as
Ti-6A1-4V. 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.
[0139] 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.
[0140] 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 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.
[0141] 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
bearing surfaces 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 bearing surfaces 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 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.
[0142] Third Embodiment
[0143] With reference to FIGS. 7 to 15 inclusive there is shown a
further preferred embodiment of the pump assembly 200.
[0144] 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.
[0145] 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
208.
[0146] 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.
[0147] 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.
[0148] 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.
[0149] 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.
[0150] In a particularly preferred form the components are welded
together.
[0151] FIG. 8 shows the impeller 204 of this embodiment and clearly
shows the support 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.
[0152] The cutaway view of FIG. 9 shows the axial cavity 219 and
also the magnet cavities 206 located within each blade 207. The
support structure 220 extending from housing cover 202 aligned with
the axis of inlet 210 and axial cavity 219 of impeller 204 is also
shown.
[0153] FIG. 10 is a side section, indicative view of the impeller
204 defining the orientations of central axis FF, top taper face DD
and bottom taper face BB, which tapers are illustrated in FIG. 11
in side section view.
[0154] 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
surface 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.
[0155] The leading edge 223 is radiused as illustrated.
[0156] FIG. 11B illustrates in cross-section the bottom edge face
222 of blade 207 cut along plane BB of FIG. 10.
[0157] The bottom face includes cap 228 utilised for sealing magnet
205 within cavity 206.
[0158] In this instance substantially the entire face 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.
[0159] The blade 207 is 6.0 mm in width excluding the radii at
either end.
[0160] 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). Whilst not preferred, patient
blood flow 236, and venous oxygen saturation 237 can be input as
well. The pump blood flow can be approximately inferred from the
motor speed and current via curve-fitted formulae.
[0161] 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.
[0162] FIG. 14 graphs pump efficiency against flow for the same
fluid over the same speed ranges as for FIG. 13.
[0163] FIG. 15 is a graph of electrical power consumption against
flow for the same fluid over the same speed ranges as for FIG.
13.
[0164] 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.
[0165] 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.
[0166] 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 7 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.
[0167] 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.
[0168] 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 support 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.
[0169] Fourth Embodiment
[0170] 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.
[0171] A further modification of this arrangement is illustrated in
FIG. 19 wherein impeller 304 includes secter-shaped blades 305
having curved leading and trailing -38 portions 306, 307
respectively thereby defining channels 308 having fluted exit
portions 309.
[0172] 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).
[0173] 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 face portions which are shaped so as to have an
increased curvature on the pressure face 351 thereof compared with
the suction face 352.
[0174] Fifth Embodiment
[0175] 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 each of which, in this instance,
comprise a series of smoothed wedges 414 with the leading edge of
one wedge directly interconnected to the trailing edge of the
preceding wedge.
[0176] 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.
[0177] It is to be understood that, whilst the example of FIG. 21
shows the surfaces of the shroud 411 angled at approximately 450 to
the vertical, other inclinations are possible extending to an
inclination of 00 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.
[0178] 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.
[0179] 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.
[0180] 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.
[0181] In such designs with thin blades, the magnets required for
the driving torque can be contained within the top or bottom
shroud-or both, along with the optional soft magnetic yokes to
increase motor efficiency.
[0182] A variation of this embodiment is to have the wedge
profiling cut into the inner surfaces of the housing and have
smooth shroud surfaces.
[0183] Sixth Embodiment
[0184] 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.
[0185] Seventh Embodiment
[0186] With reference to FIG. 26 there is shown an arrangement
having facing deformed surfaces. The rotor blade 610 includes a
deformed surface 612 at a working face thereof. In this instance
the deformation comprises curved edge 613. Relative movement of the
rotor blade 610 in the direction of the arrow with respect to
deformed facing surface 611 forming part of the pump housing (not
shown) causes relative hydrodynamic thrust therebetween.
[0187] 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.
[0188] Principles of Operation
[0189] 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.
[0190] 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.
[0191] 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".
[0192] 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
therebetween 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.
[0193] 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.
[0194] The inset of FIG. 27 illustrates conceptually how the first
deformed surface 603 may form only part of the first face 604.
[0195] 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.
[0196] 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.
[0197] 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.
[0198] 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.
[0199] 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 support and angled at an acute angle relative
to the plane of the first gap 609.
[0200] 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.
[0201] 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.
[0202] In this state the impeller 600 is effectively suspended
exclusively by hydrodynamic thrust faces.
[0203] Characteristics and advantages which flow from the
arrangement described above and with reference to the embodiments
includes.
[0204] 1. Low running speed, hence low haemolysis and controlled
fluid dynamics (especially shear stress) in the gap between the
casing and impeller. This in turn can lead to the selection of
radial off-flow and minimal incidence at on-flow to the rotor;
[0205] 2. Radial or near-radial off-flow from the impeller can be
chosen in order to yield a "flat" pump characteristic (HQ)
curve.
[0206] Control System--Detailed Description
[0207] Embodiments of the present invention relate to a non-contact
estimation and control system usable, although not exclusively,
with blood PUMP systems of the type illustrated in FIG. 28.
[0208] In this instance the estimation and control system 10
operates on and receives sensor feedback from pump assembly adapted
for implantation in human body 12 and arranged to operate in
parallel across at least a part of heart 13 so as to at least
assist if not fully take over the pumping function of heart 13.
[0209] The pump assembly 11 includes an impeller 14 having vanes 15
which, when urged to rotate by a magnetic field generated in one or
more of coils 16, 17 generates a pressure head H across the pump
assembly 11 and causes a flow of blood Q therethrough. In this
instance the impeller 14 is both a radial pump impeller and a rotor
of motor 18 by virtue of the inclusion of magnets (not shown)
within at least part of the impeller 14.
[0210] Monitoring means 19 is adapted to sense electric current
appearing in one or more of coils 16, 17 via sensing line 35 which,
in conjunction with monitoring of voltage derived from commutation
controller 32 (which injects current into one or more of the same
coils 16, 17) permits the monitoring means 19 to derive power input
(P.sub.in) consumed by motor 18 and actual rate of rotation of the
motor/impeller 14 (n.sub.a).
[0211] By means of equation 1.1 (in FIG. 28) it is thereby possible
for monitoring means 19 to calculate an estimation of flow Q
(and/or head H) for input into microprocessor 20. Microprocessor 20
accepts these estimates and, together with other desired set points
and predetermined values calculates a desired set motor speed
n.sub.set which commutation controller 32 accepts via line 33. The
commutation controller 32 then injects current into one or more of
coils 16, 17 in order to cause impeller 14 to rotate at that set
(desired) speed.
[0212] FIG. 20 illustrates the impeller 14 utilised in example 1
(to follow) in greater detail.
[0213] FIG. 3 illustrates the head versus flow characteristic
achievable with the impeller of FIG. 20 for a number of different
motor powers (Pin). FIG. 29 illustrates the characteristic curve
used by the monitoring means 19 for example 1 (to follow) in
accordance with the equation 1.1.
EXAMPLE 1
[0214] Flow rate and pressure difference (or head) are key
variables needed in the control of implantable rotary blood pumps.
However, use of invasive flow and/or pressure probes can decrease
reliability and increase system power consumption and expense. For
given fluid viscosity, the flow state is determined by any two of
the four pump variables: flow, pressure difference, speed and
electromagnetic torque (apart from the possibility of non
uniqueness of solutions). Instead of torque, motor current or input
power can be used. Thus if viscosity is known, or if its influence
is sufficiently small, flow rate and pressure difference can be
estimated from the motor speed and input power, which can be
determined from current and voltage measurements on the motor input
leads.
[0215] The centrifugal blood pumps of previously described
embodiments use a hydrodynamic bearing and can be constructed so
that the variation with viscosity is sufficiently small to enable
flow and pressure difference estimation using signals derived from
the coils 16, 17.
[0216] For this example a flow loop was set up consisting of the
pump and 2.4 m of 3/8" tubing giving a net fluid volume of 177
ml.
[0217] The fluid filled tubing was sunk into a water bath with a
controlled heater. Temperature sensors were attached to the tubing
to provide visual feedback on fluid temperature. Pressure taps were
made on the inlet and outlet nozzles of the pump which interfaced
to a differential pressure transducer with digital display to
measure pressure across the pump. A Clamp on Transonics flow probe
and meter were used to measure flow rate and input power (motor
supply voltage and current) was monitored via digital panel meters
on the power supply. Pressure was varied by adjustment of a tubing
clamp and motor speed by wuitable electrical adjustment.
[0218] Two tests were conducted. The first with 5% saline, the
second with red blood cell suspensions, haematocrit being 32%. In
both cases the circulating fluids were heated to 370C. 5% saline
was chosen since its viscosity is about that of water at 23deg
C.
[0219] Flow rate, pressure head, pump speed and electrical input
power were measured for both fluids.
[0220] Data for saline and blood was combined and correlated on a
surface plot describing both flow rate as a function of motor speed
and input power as illustrated in FIG. 29.
[0221] Curve fitting of this plot produced the equation
Q=20.29+4.731n(Pin)-55{square root}(n) where Q is flow rate in
L/min, Pin is electrical input power to the motor in Watts and n is
motor speed in rpm. The maximum error for this prediction was 4%
for the combined data. Pressure head across the pump was described
by the relationship .DELTA.P=-13.68-6.591n(Pin)+2.18e-5 (n).sup.2
with equivalent accuracy. Two different rotor designs have been
tested in this manner to date both yielding similar accuracy curve
fits of the form Q=a+b.ln(Pin)+c.{square root}(n) and of the form
.DELTA.P=a+b.ln(Pin)+c.(n).sup.2.
[0222] The viscosity of saline is approximately 1 mPas. The
Viscosity of blood (Hct=32%) given pump shear rates of greater than
100 s.sup.-1 is near 3 mPas. Blood viscosity varies from
approximately 2.4 to 4.5 mPas over the physiological range in
question for shear rates greater than 100 s.sup.-1. The variation
in viscosity from 1 to 3 mPas produced a maximum error of 4% in the
prediction of flow rate.
[0223] The pump of FIGS. 1-6 has characteristics such that the
model for flow rate prediction based on motor input power and speed
is not greatly affected by variation in viscosity. This suggests
for this pump it is possible to determine flow with acceptable
accuracy without using a separate flow sensor.
[0224] The reasons for low error in prediction given change in
viscosity are postulated as follows: Firstly that the "flat" H-Q
curves for this pump give small variation in pressure head for
given flow rates. Secondly the nature of the hydro-dynamic hearing.
Although the pump has relatively high disc friction forces, which
tend to be most sensitive to viscosity changes, the rotor in this
case conserves energy by repositioning in free space according to
the fluid viscosity. Thirdly, the size, where surface roughness is
relatively smaller than for smaller higher speed pumps. Fourthly,
allowing speed to vary around a set point due to choosing a
comparatively long time constant.
[0225] FIG. 30 illustrates the pump assembly 11 in cross section as
utilised with example 1.
[0226] FIG. 31 illustrates in cross section the coil and magnet
arrangement used in conjunction with example 1.
[0227] With reference to FIGS. 20, 30 and 31 iron yokes are placed
outside the coils to increase the magnetic flux and hence increase
motor efficiency, and also to reduce stray magnetic fields in the
body. The yokes are positioned so that the axial magnetic force on
the impeller is zero when it is central in the housing cavity.
Furthermore, the yokes are placed at considerable distances from
the impeller to keep the negative magnetic stiffness sufficiently
low that is places only a small additional demand on the
hydrodynamic suspension when the impeller shifts away from the
cavity mid-position.
[0228] Given the large distance to the yokes, a slotless winding
and axisymmetric yokes were chosen. The use of axisymmetric yokes
implies zero "cogging" torque. The winding topology coil chosen is
of "second harmonic" type with just three coils, one per phase, in
each of the body and cover windings. FIG. 31 depicts the cover
winding. The body windings align axially with the cover windings
but must be bent in several directions to avoid the volute and
inlet. This second harmonic topology avoids coil overlaps and is
consequently neat and compact and gives low copper mass. However,
it is less efficient than other winding options with greater coil
mass.
[0229] The efficiency is increased by tilting the magnet alignment
to an angle of 22.5.degree. from the pump axis (as indicated in
FIG. 30 by the magnet hatching), intermediate between the
45.degree. conical body and the flat cover. The cover coil and
axial flux form an axial flux motor, and the body coil and flux are
intermediate between an axial and radial flux motor.
[0230] The motor can be driven by a six-step, sensorless
commutation inverter. Superimposed over the coils in FIG. 31 are
magnets at an instant when the currents are switched from phases a
and c conducting to phases b and c conducting (or v.v.). Parallel
coil connection of the cover and body coils (each connected in star
configuration) enables some redundancy, in that the motor still
runs with the loss of a coil.
[0231] The materials used were Ti-6A1-4V for the housing and
impeller shell, high remanence NdFeB magnets (VACODYM 510 HR)
embedded in the impeller, iron for the yokes (mild steel in
prototypes but to be laminated silicon steel) and varnished copper
wire for the coils.
[0232] The measured negative magnetic stiffness of the teardrop
impeller is -4000 N/m (.+-.10%). The axial clearance gaps are 0.1
mm when the impeller is central (this is to match a 0.05 mm taper
on the blades for thrust generation so that after a shift of 0.05
mm, the thrust forces are maximal from one impeller face and
negligibly small from the other face). Thus if the impeller is
shifted axially by the full amount possible (as at start-up if axis
vertical), then the magnetic force on the impeller is 0.4 N force.
This is less than the impeller weight of 46 gforce, and is
considered acceptable. If the yokes were any closer, the force
would be higher, increasing the risk of touchdown. Similarly, if
the clearance gaps are increased to slacken manufacturing
tolerances, then the maximal magnetic force can be increased.
[0233] The measured motor efficiency is between 45% and 48% curves,
for speeds between 2000 rpm and 2500 rpm and motor output power
between 3 and 7W. For example, at 2250 rpm and 3 W motor output
(roughly rated conditions), the copper loss was 1.7W, the eddy loss
in the titanium was 1.0 W. and the iron loss in mild steel yokes
was 0.7 W, giving a motor efficiency of 47%.
[0234] With reference to FIG. 32 the example 1 can be applied to
the preferred embodiment of FIG. 7 to 15 comprising pump assembly
200 incorporating an estimating and control system of the type
described with reference to FIGS. 28 to 31.
[0235] 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.
[0236] 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
208.
[0237] 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.
[0238] 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.
[0239] The windings 212 and 209 are supplied from the electronic
controller of FIG. 32. Otherwise the structure is as described with
reference to the third embodiment.
[0240] Further Embodiments
[0241] 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 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.
[0242] In the examples thus far the bottom faces of the blades 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.
[0243] 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 support 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.
SUMMARY OF OPERATION PRINCIPLES
[0244] The estimation and control system described with reference
to Example 1 and the previous embodiments is "sensorless" in that
it derives an estimate of relevant pump parameters from signals
available from one or more of the drive coils of the motor. Hence
no separate sensor device is required to control the pump assembly
in use.
[0245] It is hypothesized that the ability to control the pump
assembly in this manner to a sufficiently good approximation
derives from shaping the impeller of the pump so that a relatively
flat head versus flow characteristic is obtained over the flow rate
range expected and/or required of the pump, in use.
[0246] It is postulated that relative radial off-flow and lack of
constraint of the fluid within the impeller derived from the
relatively low number of impeller blades aids in achieving the
relatively flat pump characteristic curves as shown for example in
FIGS. 3 and 13.
[0247] It is also postulated that, in the embodiments described in
the specification, the impeller blades are arranged to guide fluid
carefully through the rotor so as to reduce re-circulation. There
are also relatively large gaps between the blades so that the fluid
is relatively poorly constrained leading to loosely constrained
flow of fluid within the pump housing.
EXAMPLE 2
[0248] With reference to FIGS. 33 to 39 a specific example of a
particularly preferred rotor, centrifugal flow pump assembly
incorporating the rotor and a control system therefor will now be
provided.
[0249] The rotor 500 of this example is illustrated in FIG. 35 and
is arranged to operate within a housing structure as previously
described in this specification with reference to FIG. 7. The rotor
500 is urged to rotate by an electromagnetic field supplied via
coil structures again as previously described with reference to
FIG. 7. The control system which maintains control over the
operation of the rotor within the housing is determined by a
non-contact estimation and control system as previously described
in this specification but further subject to an optimal pumping
condition strategy as will be described below.
[0250] With particular reference to FIG. 35 it will be noted that
the impeller blades 501 are held in mechanical relationship with
each other by struts 502.
[0251] By increasing the smallest radius from the centreline to the
blades (i.e. to the nose of the blades) at the top and not at the
bottom of the impeller, an axial thrust force can be imposed on the
impeller toward the bottom. This arrangement can be carefully
designed so as to bias the load to the bottom bearing and relieve
the top bearing which is more highly loaded (in that it must resist
both axial and radial loads).
[0252] Operation Region for the Pump
[0253] With reference to FIG. 34 the pump of this example is
arranged to follow an HQ curve that does not roll-off towards
shut-off. That is, if the pressure head (H) developed by the pump
at any given operating speed (N) is plotted against flow rate
delivered (Q), then at low flow rates (and even at zero flow) there
is no loss of head compared with the head developed at the nominal
operating point. Typically in other centrifugal pumps of the prior
art pressure head developed increases with increasing flow rate
from zero or "shut-off" to a point of inflection in the HQ curve
then head reduces with further increases in flow rate. It is normal
practice in the prior art to operate a pump to the right of the
inflection point to avoid instabilities known as surge which occur
because at a given pump speed a pressure head required might be met
by one of two flow rates delivered by the pump--one at either side
of the inflection point. To the right of the inflection point,
typically the HQ curve falls steeply.
[0254] In the pump of this example, since there is no inflection
point in the HQ curve, the pump can be operated stably throughout
its entire range of flow rates. This means that the pump is
operating in the flattest part of the HQ curve and enables better
prediction of flow and pressure from parameters which may be
attained readily from motor performance characteristics (viz.:
Voltage, current and speed).
[0255] Factors which contribute to the flat HQ curve of the pump of
this example, with an absence of an inflection point, include
near-radial off-flow from the impeller, low specific-speed design
of the pump and a low number of impeller blades.
[0256] An optimal control strategy will now be described with
reference to FIGS. 33 to 39.
[0257] Optimal Control Strategy
[0258] It is the aim of the rotary blood pump and its associated
control system of Example 2 to restore normal cardiac output levels
such that the demand for perfusion is supplied by pumping as much
blood from the ventricle as is returned to it from the lungs.
[0259] Rate responsive control of the pump is described in this
example to determine the optimum point for unloading the heart
while at the same time avoiding over pumping leading to suction or
under pumping leading to regurgitation during the varying
physiological climate of every day life.
[0260] Since the pump has no valves, there is a possibility of back
flow when the pump speed is low. FIG. 36 shows the normal direction
for flow of blood through the pump from the left ventricle to
descending aorta. This point changes with preload (left ventricle
pressure) and afterload (arterial pressure) across the pump.
Furthermore, as pump speed is increased the aortic valve will
eventually remain closed and additional increases in speed will
cause collapse of the ventricle.
[0261] The rotary blood pump is sensitive to pre load and after
load if the pump speed set point has no feed back. Instantaneously
increasing the pressure head across the pump will cause an increase
in impeller speed and decrease in electrical input power and pump
flow rate. Decreasing it will have opposite effects.
[0262] In this example the time constant of the control system is
set to be relatively slow to the extent that disturbances induced
in the speed of blood flow by the action of the heart will be
counteracted by the control system relatively slowly. The resulting
variation in speed of the impeller, in use, is then utilised to
calculate an estimate of the operating point to an improved level
of accuracy.
[0263] The long time constant means that instantaneous pump speed
and electrical input power will vary cyclically under the influence
of the pumping action of the heart or, in other words, will be
modulated by the heart beat.
[0264] In this example the time constant of the control system is
set to be greater than the rotational inertial time constant of the
impeller. Specifically, in this example, the time constant is set
at 5 seconds which is longer than one cardiac cycle.
[0265] Optimal Pumping and Avoiding Over Pumping
[0266] If pump speed is set such that maximal unloading of the
ventricle is achieved and venous return is reduced as in the case
from exercise to resting, over pumping from the ventricle will
result in suction and collapse of the ventricle may occur.
[0267] As the pump speed is increased the ventricle empties and the
pressure in the ventricle during systole decreases. This is shown
in FIG. 37 as dP is reduced from dp1 to dP3 during systole. If the
amount of emptying of blood from the ventricle matches the amount
of filling, the pump is on the verge of producing suction or
negative pressures in the ventricle. Beyond the point that the
aortic valve remains closed and the peak left ventricular pressure
during systole continues to decrease and suction will begin to
occur during the diastolic phase.
[0268] Further increases in pump speed will cause the peak left
ventricular pressure to become so low that the ventricle walls will
occlude blood flow through the inlet cannula over the entire
cardiac cycle, even during systole. Suction should be avoided even
during diastole. The optimum point of pumping is just allowing the
aortic valve to open. Over pumping is considered increasing pump
flow beyond this point.
[0269] The solution to detection of the point of optimal pumping in
this example lies in the time domain.
[0270] The point at which the aortic valve just remains closed is
the point of total assist given the name OCA (optimal cardiac
assistance) This is the point at which minimum head pressure across
the pump begins to rise with increasing pump speed. In other words
during systole the left ventricle peak pressure begins to decrease
as average pump speed is increased.
[0271] Therefore for a given preload, afterload and contractile
strength of the ventricle there will be a point where optimum
unloading of the ventricle occurs. Increase in pump speed beyond
this point will result in collapse of the ventricle. This minimum
pressure across the pump during systole will produce a maximum flow
through the pump, maximum torque on the impeller and minimum
instantaneous speed.
[0272] Therefore pumping at the point of optimal cardiac assistance
and avoiding over pumping, the control algorithm should maintain
minimum pump speed such that the minimum head pressure across the
pump does not increase. Therefore the new desired set point Nnew to
hold the optimal cardiac assistance point can he defined by the old
speed value Nold reduced by a factor proportional to the increase
in minimum systolic head pressure (.DELTA.Hsys) beyond the minimum
possible head pressure (.DELTA.Hmin) Kp is the proportional
constant. This is described by equation 1.
Nnew=Nold-[Kp*(.DELTA.Hsys-.DELTA.Hmin)] equation 1
[0273] The instantaneous head pressure can be estimated by
non-contact methods as previously described in this specification
with reference to Example 1.
[0274] This is a simple control equation that can be readily
implemented in an embedded microcontroller system.
[0275] Avoiding Under Pumping
[0276] The other boundary condition of under pumping occurs when
flows through the pump become negative with diastole. Regurgitation
can cause stagnation of blood and lead to thrombus formation as
well as increasing atrial pressures leading to pulmonary
adaema.
[0277] Regurgitant or negative flow in the pump begins to occur as
pump set point speed is decreased to the extent where levels and
phase lags between pump outlet and inlet pressures during diastole
cause flow reversal.
Nregurg=N(t) for Qdiastole=0L/min equation 2
[0278] where N(t) is the instantaneous impeller rotational speed,
Qdiastole is the minimum flow rate through the pump during diastole
and Nregurg is the minimum speed at which non regurgitant flow
occurs. Flow rate can be estimated by non-contact methods as
previously described as a function of motor speed and input power
as shown earlier with reference to Example 1.
[0279] SUMMARY OF OPERATIONAL FORCES EXPERIENCED
[0280] In practical operation of Example 2 the rotor 500 should be
made to operate such that blood flow is adequate in accordance with
the constraints and the optimal control strategy described above.
In addition, whilst in operation, the rotor 500 ideally should
never make contact with the inside walls of the housing in which it
rotates. Should such contact be made then the control system should
be able to recover from this condition so as to return the rotor to
an operational condition and, in addition, damage sustained during
a touchdown must be minimised so that, upon return to normal
operation after touchdown, there is no effect on steady state
operation.
[0281] Touchdown is countered by ensuring that there is sufficient
restoring hydrodynamic force exerted upon the rotor 500 to
counteract any disturbing force experienced by the rotor 500 such
that probability of touchdown is reduced to a sufficiently low
value.
[0282] Broadly, for the centrifugal pump structure described with
reference to this example and having a rotor 500 of the type
described with reference to this example it has been found that
worst case restoring forces occur when the rotor is rotating in a
low viscosity medium and running at its lowest speed. For example,
running at approximately 1800 rpm in a blood substitute
representing the lowest viscosity likely to be encountered in
practice the axial restoring force available is approximately 2
Newton. The corresponding radial restoring force is approximately
0.5 Newton under these conditions.
[0283] In a more usual and expected operating environment with
blood viscosity of the order of 2.5 mPs and with the rotor running
at approximately 3000 rpm the axial restoring force available is
approximately 9 Newton whilst the radial restoring force is
approximately 2.25 Newton. If the speed is reduced to around 2400
rpm then the axial restoring force is approximately 5.3 Newton and
the radial restoring force is approximately 1.3 Newton.
[0284] An expected typical steady state disturbing force can be of
the order of 0.45 Newton including the effects of gravity upon
rotor 500. Magnetic field disturbances arising from the drive
mechanism can add a further 0.1 Newton of disturbing forces.
Allowing for acceleration effects on the entire assembly in vivo
and in use it is expected that typical maximum disturbing forces
encountered by rotor 500 will be of the order of 1 Newton.
[0285] Use of a shrouded rotor as described earlier in the
specification can double the radial resistance force available to
counteract disturbing forces.
[0286] In addition appropriate coatings and/or structure materials
will be used on the respective rotor 500 and at least inner walls
of the housing to minimise damage and/or damaging effects arising
from a touchdown.
[0287] Coatings and/or inherent structural materials will be
selected so as to reduce the friction co-efficient between the
rotor and the casing and also to reduce specific damage such as
gouging.
[0288] Current particularly preferred materials for this purpose
include amorphous carbon based materials or microcrystalline carbon
based materials and titanium nitride. In a particular form it has
been found advantageous to have different materials on opposing
surfaces such as, for example, titanium nitride on one of the
surfaces and the carbon based materials on the other.
[0289] Carbon based material against carbon based material
corresponding surfaces has been found to give a very low
co-efficient of friction (typically 0.05) and low damage.
Conversely, titanium against titanium has been found to give the
reverse effect and is not recommended.
[0290] Further particular preferred coating arrangements are as
follows.
[0291] Application of coatings to the Ti-6A1-4V substrate of a
blood pump.
[0292] These coatings specifically include carbon with a graphitic
microcrystalline structure, prepared using unbalanced magnetron
sputter deposition and amorphous carbon coatings.
[0293] These coatings provide a biocompatible lining for the
device, with enhanced hardness, high elastic recovery under impact
conditions, low friction coefficients of <0.06 under lubricated
conditions, and high wear resistance.
[0294] Also coatings applied using plasma immersion ion implantion
of nitrogen, titanium nitride and carbon, or combinations of these
treatments, for enhancement of hardness, improved elastic recovery
under impact conditions, low friction and high wear resistance.
[0295] Other potential candidates include pyrolitic carbon,
illumina, zirconia or combinations thereof.
[0296] Overall the desirable characteristics to be achieved by the
coatings are.
[0297] 1. Zero eddy current loss in the casing;
[0298] 2. A hard material on at least one surface to give a good
bearing property; and
[0299] 3. The materials must be biocompatible, particularly in
relation to blood contact.
[0300] More generally in relation to modification to surfaces
attention can be paid to other characteristics of the surface
structure of both the rotor 500 and the at least inner wall of the
housing with a view to providing flexibility or other dynamic
characteristics which can aid hydrodynamic bearing behaviour.
Elasto-hydrodynamic bearings are one such arrangement which can be
achieved by further attention to the materials of which the rotor
and/or housing inner walls are constructed as will now be
described.
[0301] Elastohydrodynamic Bearings (EHD Bearings)
[0302] Elastohydrodynamic (EHD) bearings rely on the principle that
if forces applied to bearings are sufficiently large then the
bearing surface will distort. This distortion can lead to a greater
efficacy of the bearings allowing greater loads to be carried for a
given bearing dimension of course the magnitude of force necessary
to distort the bearing surface must be viewed relative to the
modulus of elasticity of the material from which the bearing
surface is manufactured. For EHD to be applicable to rotary blood
pumps, the modulus of elasticity of the bearing surfaces must be
low in order for them to be distorted by forces of a few Newtons
magnitude. For these reasons polymeric materials such as
polyurethane or silicone may prove acceptable materials from which
to fabricate the bearing surfaces (these materials should sit on a
harder substrate of, say, titanium or a ceramic). The fundamental
shape of the hydrodynamic bearing, including particularly the
"deformed surfaces" would need to remain substantially the same as
for the hydrodynamic bearings previously described in this
application for use in a blood pump.
[0303] BIO-EHD Bearings
[0304] An alternative approach is to allow tissue overgrowing the
bearing substrate to act as the EHD component. The substrate may be
porous such that it allows a pseudoneointimal cell lining to grow
into the pores on the substrate surface. It is commonly reported
that the pseudoneointimal lining thickness is stable and around one
cell deep. The advantage of using a "bio-EHD" component on the
surface is that damage to the EHD component may regenerate within a
few hours of damage occurring. Possible drawbacks may be a tendency
for a bio-EHD component to sustain damage under relatively low
shear stresses commonly seen in a rotary blood pump and for
sections of bio-EHD to be stripped away forming potentially
dangerous embolii. This may be countered by additional surface
treatments which promote stability of the pseudoneointima. Once
again, the fundamental shape of the bearing, that is the "deformed
surfaces" should remain substantially the same as for the
hydrodynamic bearings of the blood pump embodiments previously
described.
[0305] The above describes only some embodiments and some examples
of a rotary blood pump and control system therefor and
modifications, obvious to those skilled in the art, can be made
thereto without departing from the scope and spirit of the present
invention.
INDUSTRIAL APPLICABILITY
[0306] The pump assembly 1, 200 is applicable to pump fluids such
as blood on a continuous basis. With its inherently simple
mechanical and control structure it is particularly applicable as
an in vivo heart assist pump.
[0307] 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 should be avoided or where leakage of the fluid should be
prevented with a very high degree of reliability--for example where
the fluid is a dangerous fluid.
* * * * *