FLUID DISPLACEMENT PUMP WITH BACKPRESSURE STOP

Description

Fluid Displacement Pump With Backpressure Stop

Technical Field

The invention relates to a fluid pump for pumping

liquid and/or gas phase materials. The fluid pump is

useful, as described in my earlier applications, in the

context of an output system of an internal combustion

engine or a turbine engine and an input system for

injecting fluid into the combustion process. The input

system, in that case, includes a displacement pump,

specifically for use with air and water, which can be

utilized as a gas compression pump in the internal

combustion engine and the turbine.

Background Of The Invention

Fluid displacement pumps are subject to a variety of

applications in engineering. For instance, such pumps are

utilized in compression systems such as air compressors

and as fluid pumps. For example, British Patent

Specification 265,659 to Bernhard discloses an internal

combustion engine with fuel pressurization separate from the combustion chamber. There, fuel is pressurized in a

compressor and the pressurized fuel is fed from the pump

to the engine through a port assembly.

U.S. Patent No. 1,287,268 to Edwards discloses a

propulsion system for a motor vehicle. There, a compressor

formed with mutually interengaging helical impellers pumps

to an internal combustion engine which is also formed with

mutually interengaging helical impellers. The internal

combustion engine drives a generator, which pumps

hydraulic fluid to individual hydraulic motors that are

disposed at each of the wheels. The impellers of Edwards

are formed with "flat" blades of a constant thickness from

the axle radially outward to their outermost tip.

The efficiency of fluid pumps with interengaging

impeller blades is dependent on the seal that is in effect

formed between the blades. While the outer seal is

relatively easily obtained with a corresponding housing

wall, the inner seal between the blades, i.e., at the

location where the blades overlap is rather difficult to

obtain. In the prior art system of Edwards, for example,

the flat blades do not sufficiently seal against one

another and the corresponding efficiency of the double impeller pump is therefore relatively low. Certain

applications of the fluid pump require a better seal and

better backflow prevention.

Summary Of The Invention

It is an object of the invention to provide a fluid

displacement pump, which overcomes the disadvantages of

the heretofore-known devices and methods of this general

type and which is further improved in terms of efficiency

and backflow prevention, and which allows essentially

continuous pumping output with negligible backflow.

With the foregoing and other objects in view there is

provided, in accordance with the invention, a fluid

displacement pump, comprising:

a housing formed with a chamber having a wall defined

by two parallel, mutually intersecting cylindrical

openings defining respective cylinder axes; and

two axles respectively disposed at and rotatably

mounted about respective axes coaxial with said cylinder

axes, said axles each carrying a helically rising blade

sealing against said wall of said housing and engaging into one another so as to form a substantially completely

closed wall within said chamber during a rotation of said

axles;

said blades having a decreasing thickness from said

axles to an outer periphery thereof.

In an alternative embodiment of the invention, the

blades increase in thickness from the axle outward.

Details of the alternative embodiment will emerge from the

following description of the figures.

In accordance with an added feature of the invention,

said blades have a rounded surface extending from said

axle to an outer periphery thereof.

In accordance with an additional feature of the

invention, said rounded surface is defined by a radius of

curvature in a radial section of said blades, said radius

being greater than a diameter of said blades. Preferably,

the radius of curvature is approximately three times the

diameter of said blades. In accordance with another feature of the invention,

said blades are conical as seen in axial section, with

mutually opposite surfaces steadily merging towards one

another from said axle to the outer periphery.

With the above and other objects in view there is

also provided, in accordance with the invention, a fluid

displacement pump, comprising:

a housing formed with a chamber having a wall defined

by two parallel, mutually intersecting cylindrical

openings defining respective cylinder axes; and

two axles respectively disposed at and rotatably

mounted about respective axes coaxial with said cylinder

axes, said axles each carrying a helically rising blade

sealing against said wall of said housing and engaging

into one another so as to form a substantially completely

closed wall within said chamber during a rotation of said

axles;

said blades having a given thickness and helically

rising along said axle with a given lead substantially

greater than the given thickness of said blades. In a preferred embodiment, the ratio of the spacing

between the blade turns (the lead minus the blade

thickness) to the thickness of the blades lies between 5/4

and 2.

The axles are preferably cylindrical, i.e., their

peripheral wall is defined by mutually parallel lines.

In accordance with an added feature of the invention,

the rounded surface is defined by a radius of curvature in

a radial section of the blades, the radius being greater

than a diameter of the blades. In a preferred embodiment,

the radius of curvature is approximately three times the

diameter of the blades.

In accordance with another feature of the invention,

the blade on each of the axles has a helical rise of

approximately 7° and the blades are substantially conical

in radial section from the axle to a periphery thereof.

In accordance with again an added feature of the

invention, the blade of one helix of the double helix are

spaced apart by a distance defined by the blades of the

other helix of the double helix. In accordance with a concomitant feature of the

invention, the blades enclose an angle of between

approximately 45° and almost 90° with the cylinder axes.

Other features which are considered as characteristic

for the invention are set forth in the appended claims.

Although the invention is illustrated and described

herein as embodied in a fluid displacement pump with

backflow stop, it is nevertheless not intended to be

limited to the details shown, since various modifications

and structural changes may be made therein without

departing from the spirit of the invention and within the

scope and range of equivalents of the claims.

The construction and method of operation of the

invention, however, together with additional objects and

advantages thereof will be best understood from the

following description of specific embodiments when read in

connection with the accompanying drawings. Brief Description Of The Drawings

Fig. 1 is a partial sectional and side-elevational

view of a fluid displacement pump according to the

invention;

Fig. 2 is a top plan view onto the impeller blades

and the housing of Fig. 1 ;

Fig. 3 is a plan view of the housing;

Fig. 4 is a plan view onto the impeller blades;

Fig. 5 is a side view of two mutually interengaging

blade structures;

Fig. 6 is an enlarged view of the detail indicated in

Fig. 5;

Fig. 7 is an axial section through the axle and a

blade of a preferred embodiment of the invention;

Fig. 8 is a diagrammatic sectional view of an

alternative embodiment of the blade structure; Fig. 9 is a diagrammatic sectional view of a further

alternative embodiment of the blade structure;

Fig. 10 is a diagrammatic section view of yet another

alternative embodiment of the blade structure;

Fig. 11 is a diagrammatic sectional view of another

alternative embodiment of the blade structure;

Fig. 12 is a diagrammatic sectional view of yet

another alternative embodiment of the blade structure;

Fig. 13 is a diagrammatic sectional view of an

alternative orientation of the blade structure;

Fig. 14 is an elevational view of two equal

orientation impeller blades prior to interengagement; and

Fig. 15 is an elevational view thereof, after the two

blades have been inserted into one another.

Description Of The Preferred Embodiments

Referring now to the figures of the drawing in detail

and first, particularly, to Fig. 1 thereof, there is seen -lo¬

an elevational view of two interengaging impellers with a

section outline of the sidewalls of a housing and a

diagrammatic view of a drive system. The fluid pump is a

double impeller system, with a first impeller 9A driven by

a first gear 14A and a second impeller 9B driven by a

second gear 14B. The impeller embodiment is a positive

displacement system and, at the same time, a back-pressure

membrane. As the ribbed impellers rotate, the fluid flow

11 (e.g., air, liquid, hydraulic fluid) is "packaged" into

chamber 30 formed between a cylindrical impeller axle 31,

a housing wall 20, and a blade 9B . Each impeller has a

respective blade 9A and 9B.

Following the helical path of the chamber 30, each

chamber formed between the turns of the blade 9B is closed

off by the blade 9A of the adjacent impeller structure.

Depending on the rotational speed of the impeller system

and the size of the chambers 30, the impellers 9A and 9B

form a pressure pump with positive displacement towards a

high-pressure chamber. The fluid flow 11 is at a lesser

pressure than in the high-pressure chamber, located above

the housing in Fig. 1. As the blades 9A and 9B of the

impeller rotate, various vertically stacked chambers are

opened and closed so that it will result in a positive flow from the bottom to the high-pressure side at the top.

At the same time, any pulsations and explosions due, for

example, to a combustion of fuel in a chamber on the high-

pressure side or any other backpressure will be prevented

from flowing back past the blades 9A and 9B . In other

words, the impeller pump is always closed with regard to a

direct backflow of the fluid out from the high-pressure

side.

The impellers 9A and 9B may be driven at variable

speed. In order to synchronize the blades 9A and 9B, they

are connected via gear wheels 14A and 14B, respectively,

connected to their axles 31. A drive 26 is

diagrammatically illustrated towards the left of the gear

14A. The drive 26 may be, for example, a gear of a toothed

rack, an electrical motor, a feedback system driven by the

output of the axles 31, or any similar controlled drive.

Any type of speed control may be implemented for the

impeller system. It is also possible, of course, the drive

the shafts 31 directly with direct drive motors. The two

spindles are engaged with the meshing gear wheels 14A and

14B. Fig. 2 is an axial plan view of the impeller system

showing the engagement or meshing of the two blades 9A and

9B and the tight placement of the impeller blades inside

the walls 20. The positive displacement force of the

impeller system is thus only slightly impaired by backflow

and leakage between the impeller blades 9A, 9B and the

walls 20 and, negligibly, between the axle 31 and the

adjacent blade 9A or 9B. The blades 9A and 9B seal tightly

against the housing wall 20. In an exemplary embodiment of

the novel fluid pump, the spacing between the outer

periphery of the blades and the inner surface of the wall

is in the range of a few mils, for example 0.1 - 0.4 mm.

Depending on its use, the fluid pump may be additionally

sealed with a silicon sealing layer provided on the inside

of the housing wall and/or on the periphery of the blades

9A and 9B .

With reference to Figs. 2 and 3, the housing of the

positive displacement system is defined by walls 20 with

rotationally symmetrical portions. In the illustrated

embodiment with the two interengaging impellers, the

housing has two intersecting circular arches that

essentially correspond to the periphery of the blades 9A

and 9B in their engagement position. A width D of the housing opening in which the impeller spindles are

rotatably disposed corresponds to a sum of the diameters

of the impeller blades 9A, 9B minus the overlap O. The

overlap 0, in turn, corresponds essentially to the rifling

depth of the impellers, i.e., the difference in the radius

of the blades 9A, 9B and the radius of the shaft 31. The

width D may also be expressed as the sum of two times the

diameter d of the shaft 31 plus two times the rifling

depth of the impellers.

As seen in Figs. 4 and 5, the blades or helical

rifling of the blades is offset by approximately 180° so

as to distribute the pumping discharge of each of the

chambers 30 into the high-pressure side. In other words,

it is advantageous for the chambers 30 to reach the top

position at which they empty into the high-pressure side

alternatingly . In the case of two blades, the offset

should thereby be in the neighborhood of 180°.

If three or more impeller spindles are used, the

housing 20 requires a corresponding modification and,

advantageously, the rotary offset of the impeller rifling

may be distributed accordingly by 360 °/n, where n is the

number of impeller spindles. The volume of the chambers 30 and the rotational

speed of the impellers defines the pump pressure and the

volume displacement per time of the impeller injection.

With reference to Fig. 6, the volume of each chamber 30

corresponds approximately to the double integral of the

differential rotary angle dθ taken through 360° and the

differential radius dr taken from the radius r of the

shaft 30 to the radius R of the impeller blade 9A, 9B,

multiplied with the blade spacing z, minus the volume

portion of the adjacent blade that engages into the space

in the center between the two spindles.

In order to maximize the seal between the blades, and

thus the seal of the backflow-preventing wall, the blades

9A and 9B are modified in terms of their curvature. In

that regard, the illustration in Figs. 1, 5, and 6 is

simplified to show the blades with a constant thickness

from the axle 31 to their peripheries. With reference to

Fig. 7, which is a sectional view taken diagonally through

the center of the axle 31 of one of the impellers, the

blades are curved from the axle outward with regard to

their thickness. The measurements and relationships among the various dimensions are best illustrated with reference

to a specific example.

In the exemplary embodiment, the blades 9 have a

diameter D = 125 mm (5 in) . The axle 31 has a diameter d =

25 mm (1 in) . The radius r of the blades, therefore, is r

= 50 mm (2 in) , measured from the periphery of the axle 31

to their outer periphery. The rise angle of the helically

winding blades 9 is about 7°. As an intermediate

production step, the blades may be tapered by a taper

angle φ = 3°. That is, the angle α formed between the

peripheral wall of the axle 31 and the blade 9 is α = 90°

+ φ = 93° at the top and at the bottom. Furthermore, the

blades 9 are curved from the inside out with a radius of

curvature R = 400 mm (16 in) . The position of the origin

of the radius R (i.e., the center of the arc) is defined

by the angle φ. For instance, if φ = 0, then the blades

are not tapered, and the origin of R lies on the

peripheral wall of the axle 31. If the blades are tapered

with φ > 0, then the origin of R is moved into the axle 31

by the appropriate amount defined by the angle φ. By

modeling the novel shape of the blades, the inventor has

been able to confirm that a proper and superior seal is

created between the interengaging impellers. Fig. 8 illustrates an alternative in which the blades

9 are only tapered with the angle cp. The surfaces are not

rounded. In a preferred embodiment of this alternative,

the angle φ = 3°.

Fig. 9 illustrates yet another alternative. Here, the

blades are not tapered, but only curved. Again, the radius

R = 400 mm (16 in) and the origin of the arc lies on the

peripheral wall of the axle 31. Accordingly, the

intersection angle α between the blade 9 and the axle 31

is α = 90°.

Fig. 10 illustrates a further variation. Here, the

inventor recognized that certain fluids (usually lower

viscosity fluids) require a less proper seal between the

blades. Accordingly, here, a spacing L between the blade

windings which defines the lead of the impeller, is less

than a thickness H of the blade 9 (note that the distance

L is not the lead of the helical winding, the lead would

be defined by the spacing L plus the height of the blade,

i.e., L+H) . Here, the difference is ΔD = L - H. The

reduction from the spacing L to the thickness H may be

from 80% to as much as 50%. In other words, a ratio L/H may range from 5/4 to 2. In the embodiments with the blade

taper and/or the curvature defined by the radius R, the

parameters L and H must be defined in dependence on the

distance r from the axle 31. That is, in that case, ΔD =

L(r) - H(r) and the spacing L and the height H of the

blade 9 is preferably chosen such that ΔD is constant.

Figs. 11 and 12 illustrate yet a further variation of

the inventive concept. In Fig. 11, the blades 9 have a

bulge in section. That is, the height H of the blade

varies from ^ at the axle 31 to H₂ at approximately half

its radial extent, and then returns to the height H_x at its

outer periphery. The embodiment of Fig. 12 is similar,

except the blade 9 thins considerably at its outer

periphery, to a height H₃ < H_x < H₂.

The embodiment illustrated in Fig. 13 provides for an

attack angle θ between the blade 9 and the axle which is

different from 90°. In a preferred embodiment, the angle θ

= 70°. It should be understood that the embodiment with

the non-orthogonal orientation of the blades, i.e., the

angle θ 90°, is not exclusive of the rounded and/or

tapered variations that are illustrated in Fig. 8, 9, 11 and 12. Further, the increased spacing ΔD illustrated in

Fig. 10 may be utilized in this embodiment as well.

It will be understood that, of a pair of blades, one

may be right -wound and the other may be left -wound. In

that case, a counter-rotation of the two blades leads to a

rise of both of the spaces 30. If the two blades are wound

in the same sense, then the blades will be rotated in the

same direction. In the former case, however, a

substantially reduced amount of friction will result

between the two sets of blades. Also, if the adjacent

blades rise in the same sense, the axes must be offset

from parallel by twice their lead angle. This illustrated

diagrammatically in Figs. 14 and 15.