Regenerative rotodynamic machines

Abstract

A regenerative rotodynamic compressor has a stator which defines, together with the periphery of an impeller, an annular flow channel of circular cross-section within which is a core ring having a circular cross-section coaxial with and of a diameter at least half the diameter of the flow channel cross-section. The impeller has a ring of curve aerodynamic blades around its periphery that project laterally into the flow channel up to the core ring. The stator, the impeller and blading, and the core ring are all made as die-cast components, the stator being divided on a stepped radial plane so as to the assemblable from two main die cast components and an insert ring, also die-cast, that in part defines the flow channel and is fitted between the two main components, being secured to one of them. Inwardly directed radial pegs on the core ring locate it in the flow channel by being clamped between the insert ring and the main stator component to which it is secured. Inlet and outlet baffles on the core ring extend into the inlet and outlet ports respectively, to guide fluid into and out of the machine with greater efficiency.

Description

TECHNICAL BACKGROUND 3

This invention relates to regenerative rotodynamic machines, and more especially to regenerative compressors and exhausters.

A regenerative or peripheral pump is a rotodynamic machine which permits a head equivalent to that of several centrifugal stages to be obtained from a single rotor with comparable tip speeds. The impeller can take the form of a disc with a set of flat vanes projecting axially at each side near the disc periphery. Around the greater portion of the periphery the vanes project into an annular channel of which the cross-sectional area is greater than that of the impeller vanes. Over a sector of limited extent between the inlet and discharge the annular channel is reduced to a close running clearance around the impeller. This sector is called the stripper seal and its function is to separate the inlet and discharge ports, thereby forcing fluid out through the discharge port. The stripper seal allows only the fluid between the impeller vanes to pass through to the inlet.

The advantage of pumps of this type lies in the generation of a high head at low flow rates. They have a very low specific speed. Although their efficiency is not very high, pumps of this type have found many applications in industry where it is preferred to use rotodynamic pumps in place of positive displacement pumps for duties requiring a high head at low flow rates. Their simplicity, the low noise levels generated and the absence of problems due to lubrication and wear, give advantages over positive displacement pumps, despite the lower efficiency.

The flat vane regenerative pump has been adapted for the compression of gas. The advantage lies in the low specific speed giving a high pressure ratio together with a low flow rate for a given size of machine. Further advantages are oil free operation and freedom from stall or surge instability.

In such a compressor, the gas follows a helical path through the annular channel and passes through the vanes a number of times in its peripheral path from the inlet port to the discharge port. Each passage through the vanes may be regarded as a stage of compression and thus the equivalent of several stages of compression can be obtained from a single impeller. However, this pumping process by means of flat vanes cannot be considered as efficient. The fluid between the vanes is thrown out and across the annular channel and violent mixing occurs, the angular momentum acquired by the fluid in its passage between the vanes being transferred to the fluid in the annular channel. The mixing process is accompanied by the production of a great deal of turbulence and this implies an undesirable waste of power.

Several early theories of the fluid-dynamic mechanism of a regenerative pump were published. These theories were reviewed and compared by Senoo (A.S.M.E. trans. Vol. 78, 1956, pp. 1901-1102). Differences occur in the assumptions made, but in principle the various theories appear to be compatible. Senoo and Inversen (A.S.M.E. Trans. Vol. 77, 1955, pp 19-28) considered turbulent friction between the moving impeller and the fluid as the primary force causing the pumping action. Wilson, Santalo and Oelrich (A.S.M.E. Trans. Vol. 77, 1955, 1303-1316) regarded the mechanism as based on a circulatory flow between the impeller and the fluid in the casing with an exchange of momentum between the fluid passing through the impeller and the fluid in the casing.

Much later, compressors with considerably better efficiency were proposed by Sixsmith ("The Theory and Design of a Regenerative Compressor" presented before The Institute of Refrigeration, May 1981) in which the previously conventional radial vanes are replaced by aerodynamic blading. The annular channel is provided with a core to assist in guiding the fluid so that it circulates through the blading with reduced loss. The core also acts as a shroud closely surrounding the blades at their tips to reduce losses due to the formation of vortices at the tips of the blades. An improved commercial compressor embodying inter alia these features is described in our EP No. 0011983 corresponding to U.S. Pat. Nos. 4,306,833 and 4,334,821.

Although the compressor of EP No. 0011983 has performed very satisfactorily, and better than its contemporaries in the field of commercial regenerative compressors, there remains room for further improvement. It is therefore a general object of the present invention to achieve a compressor of this class that has better performance while at the same time being easier and less expensive to manufacture.

Prior to the present invention, workers in the art of designing these aerodynamically-bladed machines were motivated by a number of preconceptions, thus:

(i) Noise in operation is always a problem and the Roots blower is an exceptionally noisy machine--therefore, design improvements to reduce noise will require a construction that is progressively less like a Roots blower;

(ii) The clearances between the moving parts are crucial to performance and must be maintained at a minimum in manufacture by expensive machining techniques.

(iii) The small clearances require special expensive precautions to be taken in choice of materials and manufacture in order to avoid the machine seizing with the considerable heat generated in operation.

(iv) High precision - fitting parts are also necessary to prevent radial leakages from the blade channel.

SUMMARY OF THE INVENTION

By the present invention, we have overturned these preconceptions and arrived at a machine which, while being surprisingly inexpensive to construct, gives a better and more reliable performance than previous machines.

According to the invention, there is provided a regenerative rotodynamic machine comprising a stator and a rotary impeller co-operating to define, adjacent the impeller periphery, a flow channel extending circumferentially to form an annulus coaxial with the impeller, the cross-section of the flow channel being circular except at the location of a stripper seal in the flow channel that separates an inlet and an outlet respectively for fluid entering and leaving the flow channel, a static annular core ring contained within the flow channel and coaxial with the flow channel and the impeller, the core ring having, except at the stripper seal, a substantially circular cross-section coaxial with the circular cross section of the flow channel, and a ring of curved aerodynamic blades on the impeller projecting into the flow channel into close proximity with the core ring, the aforesaid components being adapted to be manufactured for assembly as die-cast parts.

BRIEF DESCRIPTION OF THE DRAWINGS

A machine embodying the invention will now be described by way of example and with reference to the accompanying drawings, in which:

FIG. 1 is a view of a compressor according to the invention, in axial section on the line 1--1 of FIG. 2,

FIG. 2 is an external elevation of the compressor in the direction of the arrow 2 of FIG. 1,

FIG. 3 is a partial view in section on the line 3--3 of FIG. 2 showing the stripper seal of the compressor,

FIG. 4 shows velocity diagrams for the impeller blades of a compressor of this kind,

FIG. 5 is a diagrammatic cross-section through the blade channel of the compressor at the region of the inlet port,

FIG. 6 is a view in section on the line 6--6 of FIG. 5,

FIG. 7A is a diagram of the blade channel in the region of the outlet port,

FIG. 7B is a diagram showing velocity distributions at the region of the outlet,

FIG. 8 is a plot of comparative curves for adiabatic efficiency,

FIG. 9 is a diagram of blade profiles, and

FIG. 10 is a diagram useful in determining the cross-section geometry for the flow channel and the core ring.

DETAILED DESCRIPTION OF THE DRAWINGS

FIG. 1 shows a regenerative rotodynamic compressor according to the invention consisting primarily of a rotary impeller 10 inside a stator or casing 11. The stator defines internally a circular impeller chamber 12 which, at one side, is in communication around its peripheral margin with an annular channel 13 of circular cross-section that is offset axially with respect to the impeller chamber 12. At the side toward which the channel 13 is offset, the side-wall 12A of the impeller chamber has a large central bore 14 in which is disposed a rotary flange 15 at the end of a drive-shaft 16, to which flange the impeller 10 is bolted by bolts 15A. The drive-shaft 16 is supported in a roller bearing 21 and ball bearings 22, 23 housed within a stationary sleeve 17 which extends axially away from the side-wall 12A and is held rigid with the casing 11 by a spider of circumferentially spaced axial webs 18 that extend both radially outwards and also axially from the side-wall 12A almost to the remote end of the sleeve 17. The outer races of the bearings 21, 22, 23 are located in a liner 19 fitted within the sleeve 17; the inner races on the shaft 16 are located by a shoulder 16A on the shaft and spacers 16B, and are retained by a nut 16C. A screwed adjustor 16D in the end of the liner 19 bears against the outer race of the roller bearing 21. The drive-shaft 16 projects beyond the end of the sleeve 17 and liner 19 and has a shaft key 24 on its external end for coupling to a drive motor (not shown).

For the purpose of manufacture and assembly, the casing 11 is split into two parts at a stepped radial plane indicated at 25. One part 11A therefore comprises the right-hand wall (as seen in FIG. 1) of the impeller chamber and a peripheral flange that provides the circumferential wall of the impeller chamber and is also internally profiled, as at 26, to provide a portion of the wall at the right-hand side of the blade channel 13. The other part 11B comprises the left-hand wall 12A of the impeller chamber, the left-hand half of the blade channel 13, the spider 18, the sleeve 17 and an integral inlet/outlet port block 28. Both these parts are designed for easy manufacture by die-casting and are secured together by bolts 31.

The remainder of the blade channel 13 is formed by a separate profiled insert ring 27, also a die-casting, that fits on to and around the step that is present on the part 11B due to the stepped plane 25 and is profiled at its outer circumference, as at 29, to provide a sector of the blade channel circular cross-section, as shown. This ring is secured in place by screws 30. There remains a gap in the blade channel circular section where it communicates with the impeller chamber 12 and through which a ring of blades 32 on the side of the impeller 10 around its peripheral margin project into the blade channel 13.

Concentric with the blade channel is a static annular core ring 33 which is of circular cross-section except for a flat 34 directed toward the impeller blades 32. The core ring is split at its maximum diameter on a plane 35 radial to the axis of the machine so that it can be readily manufactured as two die-castings and it is largely hollow, to reduce weight, as indicated as 33A, but with blocked portions 37 provided at circumferential intervals to receive screws 36 for securing the two parts together. At the inner circumference of the core ring there are spaced pegs 38 projecting radially inward from the blocked portions 37 whereby the core ring is secured in place, these pegs being clamped by the profiled insert ring 27. At the region of the inlet/outlet ports there is a gap in the core ring 33 which is occupied by a flat wall formation 39 of the part 11B of the casing (FIG. 3) that reduces the blade channel 13 locally to half its cross-section. The other part of the core ring at this region is formed to provide an integral stripper block 40 which, as can be seen, is shaped to fill the remaining half of the blade channel except for the particular area in which the impeller blades 32 run. The stripper block 40 is screwed to wall 39 by screws 40A to anchor both the block and the core ring.

The impeller 10 and the blades 32 are designed for die-casting as one integral part. The opposite faces of the impeller are `scalloped` with sectoral bands 41 alternating with sectoral recesses 42; this removes mass while giving high resistance to distortion under heat. The removal of mass in this way is highly beneficial in that the machine can be started by its synchronous driving motor without the need for either a separate high starting torque starter motor or an unloading device to relieve fluid pressure on starting, as typically used in the case of a Roots type positive displacement machine.

The blades on the impeller are curved in the radial plane of the impeller, being concave and convex, respectively, at their leading and trailing surfaces, and they are aerodynamic in that they have aerofoil profiles in which the thickness of the blade varies across its chord. It can also be seen in FIG. 1 that the chord of the blades narrows toward their extremities, so that at the blade extremities the chord dimension matches that of the flat 34 on the core ring. Velocity diagrams for the impeller blades are given in FIG. 4, the blades here being shown stylised and not in actual profile. In the diagrams, the angle .beta..sub.1 +.beta..sub.2 through which the fluid is turned at each blade is shown as 90.degree.; the angle should be at least this and the best angle appears to be 90.25.degree.. Such blades are intended to be driven at a nominal speed of 3000 r.p.m. by a synchronous electric motor. Hitherto, there was, to our knowledge, no published method of determining the number of blades to employ, but we have discovered the optimum, and the limit, for the number of blades or vanes, to be given by the following empirical formulae: ##EQU1## where B=number of blades

R=mid-blade radius (m) Q=flow-rate (m.sup.3 /min) N=rotational speed (rev/min) and in any event, the number should be less than: ##EQU2##

To reduce loss caused by the fluid angle at entry to the blading varying (i.e. incidence loss) the correct rounded blade profile should be achieved. By accurately controlling the blade geometry at the die-casting stage one can substantially reduce the need to machine afterwards to ensure the correct profile.

FIGS. 5 and 6 are diagrams of the inlet port. A baffle plate 44 is provided to guide the fluid entering the inlet 43 more efficiently into the machine and this gives a significant improvement in adiabatic efficiency. This baffle extends from the radially inward surface of the core ring 33 out into the inlet 43, where it lies against the adjacent side surface of the inlet, and it can be cast integral with the core ring.

FIG. 7A is a diagram of the blade channel at the region of the outlet port 45. In that diagram, the exit angle .phi. is the angle between the axis of the outlet passage 45A and the radial line through the rotational axis of the machine that passes through the intersection point of the outlet passage axis and the circle that defines the centre axis of the annular flow channel. We have discovered that an exit angle .phi. of 50.degree.-90.degree., preferably 70.degree., is beneficial for high pressure ratios across the machine and, in addition to setting the exit passage of the outlet port at the appropriate angle, a baffle 46 may be provided, ahead of the stripper seal 40, to give the correct angle to discharge the fluid most efficiently into the outlet port. Again this baffle can be cast integral with the core ring.

The conditions at exit are quite important to obtaining good efficiency and the precise exit angle employed will depend on other design parameters of the machine, particularly pressure ratio and rotational speed, since it is possible for a reverse air flow to be set up in the region of the exit in the general direction indicated by the arrow F. FIG. 7B illustrates the effect of changes in rotational speed upon the velocity distribution across the flow channel. As can be seen, there may be a negative velocity or reverse flow in the radially inner region of the flow channel at low speeds.

Referring again to FIG. 1, the sector faces of the `scalloped` impeller disc 10 next the casing wall 12A give a radial seal quite adequate to maintain the machine efficiency without the need for providing any special additional machined sealing plates. At the opposite side of the impeller, a labyrinth type seal 47 is provided around the impeller mounting flange 15 to cooperate with the inner wall of the bore 14. However, in the case of handling expensive or toxic gases a completely gas-tight form of seal can readily be provided at this point.

As will be seen the machine is designed to minimise heat transfer to the shaft bearings, with air space 48 between the bearing sleeve 17 and the machine casing and impeller. If desired, stirrer blades can be cast on the back of the impeller mounting flange 15 to promote the circulation of cooling air.

Not only does the machine described represent a significant leap forward in performance but also, as the principal parts are made by die-casting in aluminium and require a minimum of machining, it is surprisingly inexpensive to produce and light in weight. FIG. 8 shows comparative curves of adiabatic efficiency plotted against pressure for different speeds of rotation. The full line curves illustrate what can be achieved with the present machine as compared with the curves shown in broken lines for the best prior art regenerative rotodynamic machine.

As previously indicated, the blades shown in FIG. 4 are stylised and do not represent the actual blade profiles employed. Suitable blade profiles are shown in FIG. 9.

Referring now to FIG. 10, the radius r.sub.2 of the cross-section of the core 33 will ordinarily be at least half the radius r.sub.1 of the cross-section of the flow channel 13, and we have found that the relationship can be advantageously determined by the following empirical formula: ##EQU3## Where .alpha.=included blade angle (radians).

R=mid-blade radius (m). r.sub.1 =open channel radius (m). r.sub.2 =core radius (m). Q=flow (m.sup.3 /hr). N=rotational speed (rev/min). K is a constant. Ideally, the constant K=1/35.

Modifications of the machine described with reference to the drawings are, of course, possible without departing from the scope of the invention. Thus, as a minor point, it is appropriate to omit the roller bearing, shown as a third shaft bearing in FIG. 1, in some machine designs. More significantly, although in the machine described the blades are on one side of the rotor peripheral margin they could instead be situated on the extreme rotor periphery, projecting into a blade channel annulus surrounding the rotor periphery. However, referring again to FIG. 10, whereas the blades 32 are there shown in the same orientation in relation to the flow channel 13 and the machine rotational axis 50 as in FIG. 1, investigations are still being carried on to determine the best orientation and there are some indications that, of the alternative blade orientations such as are indicated at 51 and 52, that at 52 with the blades projecting from the impeller periphery radially inward in relation to the machine rotational axis 50 may be the best.

Claims

1. A regenerative rotodynamic machine comprising a stator and a rotary impeller co-operating to define, adjacent the impeller periphery, a flow channel extending circumferentially to form an annulus coaxial with the impeller, the cross-section of the flow channel being circular except at the location of a stripper seal in the flow channel that separates an inlet and an outlet respectively for fluid entering and leaving the flow channel, a static annular core ring contained within the flow channel and coaxial with the flow channel and the impeller, the core ring having, except at the stripper seal, a substantially circular cross-section coaxial with the circular cross section of the flow channel, and a ring of curved aerodynamic blades on the impeller projecting into the flow channel into close proximity with the core ring, the aforesaid components being adapted to be manufactured for assembly as die-cast parts.

2. A machine according to claim 1, wherein the diameter of the core ring cross section is at least half the diameter of the flow channel cross-section.

3. A machine according to claim 1 or claim 2, wherein the radii of the cross-sections of the core ring and the flow channel are defined by the relationship: ##EQU4## where .alpha.=included blade angle in radians (FIG. 10)

R=mid-blade radius (m)

r.sub.1 =flow channel radius (m)

r.sub.2 =core ring radius (m)

Q=flow rate (m.sup.3 /hr

N=rotational speed (m.sup.3 /hr)

K is a constant, preferably 1/35.

4. A machine according to claim 1 or claim 2 wherein the core ring has a flat formed on it facing the ring of blades and the blade extremities lie in a plane parallel and close to the flat.

5. A machine according to claim 1, wherein the number of blades B in the ring of blades is less than: ##EQU5## and is preferably given by: ##EQU6## where R is the mid blade radius (m)

Q is the flow rate (m.sup.3 /min)

N is the rotational speed (rev/min)

6. A machine according to claim 1, wherein the impeller and the ring of blades are made as one integral die-casting, and the stator is made principally as two die-cast components that are secured to one another on a generally radial plane of division.

7. A machine according to claim 6, wherein the stator plane of division is stepped axially in the region of the flow channel, the flow channel periphery being defined in substantial part by said two principal die-cast stator components and as to a further part by a separate profiled die-cast insert ring that is fitted between said two principal components and is secured to one of them.

8. A machine according to claim 1 or 2, wherein the core ring is split on a radial plane into two rings each of substantially semi-circular cross-section that are separately die-cast and secured together, the core ring further having pegs or spigots at circumferential intervals whereby it is located in the flow channel.

9. A machine according to claim 8, wherein the pegs on the core ring extend radially inward and are clamped between said insert ring and the principal stator component to which it is secured.

10. A machine according to claim 9, wherein the stripper seal is a die cast part of the core ring and is secured to the principal stator component to which the insert ring is secured.

11. A machine according to claim 7 or claim 9 or claim 10, wherein the impeller is secured on a shaft mounted in bearings inside an axially-extending sleeve that is integrally die-cast with the principal stator component to which the insert ring is secured.

12. A machine according to claim 11, wherein an inlet baffle is provided extending from the core ring into the inlet port to guide fluid more efficiently into the machine.

13. A machine according to any claim 12, wherein an outlet baffle is provided extending from the core ring into the outlet port to guide fluid more efficiently out of the machine.

14. A machine according to claim 1 or claim 6, wherein the blades are radially outward of the core ring and are disposed on the impeller to project inwardly therefrom toward the rotational axis of the machine.