WO2002012723A1

WO2002012723A1 - Multi-stage dry vacuum pump

Info

Publication number: WO2002012723A1
Application number: PCT/IB2001/001412
Authority: WO
Inventors: Alan Paul Troup; Linda Troup
Original assignee: Alan Paul Troup; Linda Troup
Priority date: 2000-08-03
Filing date: 2001-08-03
Publication date: 2002-02-14
Also published as: GB0018928D0

Abstract

A dry vacuum pump consisting of at least two stages of pumping disposed in series, in which a rotor (20) contains two cylinder bores (20a, 20b) within which two piston assemblies (31) and (32) are caused to reciprocate by means of a crank mechanism. The cranks are positioned inside the pistons and operate within slots which pierce the body of each piston assembly. A static crankshaft (27) extends the length of the rotor (20) and is internal to the rotor. The shaft passes through the piston assemblies via the slots. The piston assemblies are thus caused to reciprocate synchronously with the rotation of the rotor in antiphase to each other. The rotor is caused to rotate within a static, hermetically-sealed stator body (30) by means of an external motor.

Description

MULTI-STAGE DRY VACUUM PUMP

BACKGROUND OF THE INVENTION

In order to exhaust a vessel of most of the gas or vapour which it contains, a vacuum pump is employed. Such pumps serve to reduce the pressure in a vessel from a starting pressure (typically atmospheric pressure) to some much lower pressure (typically in the range 0.001 mbar to 10 mbar) . Vacuum pumps commonly employ an oil as both a lubricant for the sliding or rolling parts of the mechanism and as a liquid seal within the mechanism. Examples are oil-sealed rotary vane pumps and rotary piston pumps. It is a major problem with such pumps that the lubricant can find its way back into the vessel ("back migration" or "backstreaming") where it can contaminate the vessel or any process being carried out within the vessel. Also, the gas or vapour, in its progress through the pumping mechanism, can become heavily contaminated by lubricant.

Over the last 15 years, vacuum pumps have been developed which use little or no oil in the parts of the mechanism in contact with the gases to be pumped (the "swept volume") - these are termed "dry" pumps although it should be noted that parts of the mechanism not in contact with the pumped gases (e.g. gear boxes) will be lubricated with oil or grease. Examples are industrial 2-shaft, multistage machines using non-contacting rotors with Roots or claw or screw profiles.

While 2-shaft dry machines are now commonly used for industrial processes, they tend to be expensive and are physically large. For general light duty (for instance, in laboratories), smaller cheaper dry machines are preferred e.g. carbon vane pumps, diaphragm pumps . Further examples of such dry pumps are :

(a) The scroll pump which causes two intermeshed spiral profiles to precess without rotation in order to achieve a multi-staged pumping action. It is a single shaft machine.

(b) The piston pump, which employs one or more reciprocating, polymer-coated pistons in sliding contact within one or more smooth metal cylinders utilising cranks or other mechanical arrangements to drive the piston (s). The various pumping stages can be connected in series or in parallel to achieve different pumping performances, (c) A variant of the dry piston pump is the diaphragm pump, which uses one or more flexible membranes to expand and contract a pumping volume. The various pumping stages can be connected in series or in parallel to achieve different pumping performances .

The degree to which a mechanism can legitimately be claimed to be dry is not precisely defined. While piston pumps and diaphragm pumps can be totally dry within the swept volume, other commercial dry pumps employ shaft seals or bearing systems where a small amount of specially formulated oil or grease is exposed to the swept volume. Such pumps are usually dry enough to satisfy the contamination requirements of most applications.

Design Requirements for a Low-cost Dry Vacuum Pump

It has long been recognised by those skilled in the art of vacuum pump design that the dry mechanism likely to have the property of lowest manufacturing cost would have many similarities to an oil-sealed rotary vane pump (OSRV) . The OSRV pump gives the best all-round vacuum performance at the lowest cost, if issues of oil contamination and degradation are ignored.

The most desirable properties for a low cost dry pump are: 1. A single rotor design, employing simple rotation.

2. The ability to mount two or more pumping stages onto the single rotor to achieve serial pumping, allowing lower pressures to be reached. Modern general purpose dry pumps should achieve an ultimate pressure of less than 1 mbar. 3. The pumping volumetric capacity (measured in m³/hr) should fall in the range 5 -30 m³/hr, the range most commonly used in general or laboratory applications. 4. The mechanism should rotate at the lowest possible speed for reasons of reliability, noise and vibration, and preferably at 4-pole induction motor speed (25Hz for Europe, 30Hz for USA) to allow low-cost motors to be used. High speed mechanisms can bring problems of bearing life, vibration and lubrication and usually require manufacturing precision which is inconsistent with lowest cost of manufacture.

5. Ideally no mechanical valves but particularly no interstage valves 6. Low wear rate of contacting surfaces in relative motion - if two components must slide against one another without oil or grease lubrication, it is important to reduce wear to a point where acceptable service life before maintenance is achieved.

7. Simple geometry of component parts to reduce machining complexity and (more importantly) difficulties in checking dimensions during manufacture. Scrolls are examples of piece parts whose geometries are extremely difficult to check.

SUMMARY OF THE INVENTION

The invention described here meets most of these criteria and is, therefore, an important new concept. The invention is a small low-cost, single shaft, multistage dry vacuum pump having the property that a small amount of oil or grease is exposed to the swept volume but is expected to be dry enough to meet the requirements of most applications . The concept is called the RotoSync vacuum pump.

The key advantages of the new RotoSync dry pump concept are: • High volumetric capacity in a small machine

• Good ultimate pressure suitable for backing modern turbopumps

• Simple mechanism involving simple rotation at low speeds

• Requires no interstage valves and has the potential to operate without an exhaust valve Three related but separate inventions are described. The first is a new concept for a mechanism for a vacuum pump. The second is a new concept for a self-aligning seal to improve the sealing between a rotor and its associated stator end-face. The third is a new concept for the interstage dynamic seal required for the RotoSync pump to achieve low pumping pressures.

The first invention is a dry vacuum pump having two or more stages in a single rotor, each stage having a solid double ended piston assembly reciprocating in respective diametral through hole of the rotor and constrained to move synchronously within the through hole which is itself in rotation with the rotor and in line with inlet port and outlet port of a surrounding stator body. According to an advantageous arrangement, the dry pump may comprise :

- a stator body with at least one cylindrical inside cavity having an axis and opposed inlet ports and outlet ports ,

- a rotor , rotating within the inside cavity of the stator body around the axis of the inside cavity , and having at least two diametral through holes , each diametral through hole having two opposed large diameter bores which form two piston cylinders , the rotor having a drive end and an opposed second end, the diametral through holes being separated by a intermediate wall having a through passage around the axis of the stator body , the second end having a through passage around the axis of the stator body ,

- two solid piston assemblies , each of them comprising two opposed pistons , slidingly mounted in the respective piston cylinders of the respective diametral through hole of the rotor ,

- a stationary crankshaft , having a proximal end which is fixed rigidly in the stator body , and having at least two cranks which are laterally offset from the axis of the inside cavity of the stator body by an amount h and which are operatively connected to the respective piston assemblies so that the piston assemblies reciprocate synchronously along the respective diametral through holes of the rotor when said rotor is rotated in the stator body and around the crankshaft , the piston assemblies have corresponding transverse through passages between the pistons thereof,

- the crankshaft crosses through said through passages of the rotor and said through passages of the piston assemblies , and the cranks are inside the corresponding through passages of the piston assemblies . Advantageous embodiments are defined by other dependant claims . BRIEF DESCRIPTION OF THE DRAWINGS

Figure 1 is a perspective view of a rotor body and two piston assemblies ; Figure 2A is a perspective view of a rotor assembly with a crank mechanism inside ;

Figures 2B and 2C are side cuts of the piston assembly of fig.2A ;

Figures 3A-3H diagrammatically illustrate the various stages of the translation of the piston assembly ;

Figures 4A-4D are side cross sectional views showing a preferred arrangement for a two stage pump ;

Figure 5 is a perspective view of a rotor assembly with the crank mechanism outside ; Figures 6A and 6B illustrate the possible leak paths ;

Figures 7A-7D show the means for controlling the end-face leaks ; and

Figures 8A and 8B show the means for controlling the interstage leaks

Description of the Principle of the RotoSync Dry Pump

Fig. 1 shows a rotor 20, shaped as a right solid cylinder, with a circular peripheral face 20j and opposed planar end faces 20k, 201. The rotor 20 is radially drilled through with two large diameter holes 20a, 20b which form two piston bores into which are inserted two sliding solid piston assemblies 31 and 32. Each piston assembly is fitted with two piston rings respectively to effect gas sealing against the walls of the bores and to form two opposed pistons 31a, 31b ; 32a and 32b. The total length of the piston assembly 31, 32 is less than the length of the hole 20a, 20b by an amount 2h, where h is defined below.

Assuming the rotor 20 is at rest, it is clear that by moving each piston assembly 31, 32 backwards and forwards along its respective diametral through hole 20a, 20b, there are two positions where the ends of the piston assembly 31, 32 can become flush with the surface of the rotor 20, provided that the ends of the pistons 31a, 31b ; 32a, 32b are shaped and aligned to match the curvature of the circular peripheral face 20j of the rotor 20. The total distance that the piston assembly 31, 32 must move to translate from one flush position to the other is 2h. Each piston assembly 31, 32 is hollowed out by machining, and is pierced through by respective through passages 33 and 34 between the pistons 31a, 31b ; 32a, 32b. Referring to Figs. 2A, 2B and 2C within the body of each piston assembly 31, 32, a crank mechanism is inserted into the through passages 33, 34 and retained by a gudgeon pin 40. The crank mechanism is described in detail below. A circular shaped crank 46 is mounted on a stationary crankshaft 27 which is fixed rigidly in space and its centreline coincides with the centreline 8 of the rotor 20 - the means of achieving this mechanical arrangement is also described below. The centre 46b of the crank 46 is offset from the centreline 8 of the crankshaft 27 by an amount h - this offset defines the quantity h. A connecting body 43 is fitted between the crank 46 and the piston assembly 31. The connecting body 43 has a first end 43a and a second end 43b. The first end 43a of the connecting body 43 is pivotably secured by the transverse gudgeon pin 40 into the crown of the piston assembly 31 and can swing through a limited angle but never far enough to allow the connecting body 43 to protrude beyond the envelope of the piston assembly 31. The second end 43b of the connecting body 43 can freely rotate around the crank 46. Fig. 2B and Fig. 2C show a cross-section of the piston assembly 31 illustrated in Fig. 2A, with the section being perpendicular to the crankshaft 27. In Fig. 2B, the cross-section reveals the crank 46 which is a slide fit on the crankshaft 27 but prevented from rotating by means of a key 48. The connecting body 43 is free to rotate about the crank 46 by means of a large diameter bearing 44. In the position illustrated, the distance from the centre of the gudgeon pin 40 to the centre of the crankshaft 27 is a maximum.

If the rotor 20 is now rotated through 90° counter- clockwise, the piston assembly 31 rotates to the position shown in Fig. 2C. Because the crankshaft 27 is fixed (and hence the crank 46 is fixed) , the connecting body 43 rotates on its bearing 44 to the position shown. It is clear that the distance from the centre of the gudgeon pin 40 to the centre of the crankshaft 27 is now shorter than in Fig. 2B and, therefore, the piston assembly is forced to translate along the through hole 20a of the rotor 20. Figs. 3A-3H diagrammatically illustrate the various stages of the translation of the piston assembly 31 in its diametral through hole 20a, with rotation in the counter-clockwise direction frozen at 45° respective increments.

The rotor 20 rotates within a circular cylindrical inside cavity 30a of a stator body 30 (partially shown) having opposed inlet port 30b and outlet port 30c. The crankshaft 27 is fixed rigidly to the stator body 30, with the centreline 8 of the crankshaft being alligned with the axis of the inside cavity 30a of the stator body 30 and with the axis of the rotor 20. The piston through hole 20a has two opposed end large diameter bores which form two end piston cylinders 20m and 20n (Fig. 3A) . The pistons 31a and 31b of the piston assembly 31 are slidingly mounted in the respective piston cylinders 20m and 20n.

The series begins at 0° (Fig. 3A) with the piston assembly 31 at the lower extreme of its travel and piston 31b flush with the peripheral face of the rotor 20 : note the corresponding adjacent edges 20i of the rotor 20 and 31i of the piston assembly 31. Because the piston assembly 31 is shorter than the diameter of the rotor 20 by an amount 2h, a volume of air (denoted VI) exists at the other end of the piston assembly 31. As the rotor 20 rotates through the 45° (Fig. 3B) , 90° (Fig. 3C) and 135° (Fig. 3D) positions, the volume VI is transported anticlockwise and is reduced in size by the movement of the piston assembly. At the 180° position (Fig. 3E) , the volume VI has been reduced to zero and the other piston 31a is flush with the periphery of the rotor 20. The edges 20i and 31i are apart from each other by 2h. Thus, in half a rotation, volume VI has been transported from the top of the figure (the "inlet") to the bottom (the "outlet") and has been compressed to minimum volume. During the second half rotation, it is V2 which is transported and compressed. This shows that the mechanism has the important property that, as the rotor 20 rotates, an inlet volume (VI or V2) is transported with compression to the outlet twice per rotation with a total stroke length of 2h.

Let the radius of the piston bore be r, expressed in millimetres. It can be shown that VI (expressed in m³) is given by:

Vl=2xl0^"s h π r

As the rotor 20 rotates, the piston assembly 31 slides such as to shrink the volume VI to a theoretical value of zero, this being attained when the end 31b of the piston assembly 31 is flush with the surface of the rotor 20. The volumetric displacement of the mechanism (based on its geometry) is therefore 2 V per revolution. If the rotor 20 rotates f times per second, the volumetric capacity S (expressed in m³/hr) is:

S=4.52 xl0^~b f h r

(f in Hz, h,r in mm, S in m³/hr) . For example, f=25 Hz, h=10mm, r=42.5 mm gives S=20.4 m³/hr.

To accommodate a large piston assembly 31 of diameter 85 mm, the rotor diameter would be approximately 140 mm, although the choice of diameter is under the designer's control. It can be seen that the mechanism yields a very large volumetric speed for a modest size of rotor, as well as acting as a positive displacement compressor with self-valving between the inlet and outlet. The self-valving action is provided by the periphery of the rotor 20 ; it can be seen in Figs. 3A-3H that the outlet 30c is always isolated from the inlet 30b by a length of the rotor periphery approximately equal to {half the rotor circumference minus the piston diameter}, ignoring the dimensions of the ports.

The Compression Ratio K is given by the ratio of the outlet pressure P_out of a stage to the inlet pressure P_in

Assuming that gas or vapour enters the volume VI at the pressure Pι_n, it will be compressed as the volume is reduced by the movement of the piston assembly 31. Clearly, the maximum value of K = K_m is limited in practice by gas properties, mechanical tolerances, mechanical clearances, etc. This value K_m, defined as the maximum achievable ratio of outlet pressure to inlet pressure, is of considerable importance if the mechanism is to be useful in a vacuum pump. K_m is achieved when the net flow of gas through the inlet (or outlet) is zero and the only gas being pumped is due to internal leakage.

Relative Motion of Piston and Rotor

Assuming an offset h of 10mm, the piston stroke is 20mm. The piston therefore travels 40mm in a single reciprocation within the rotor cylinder. If the rotor rotates 25 times per second, the total distance travelled by the piston in one second is 1 metre and the time-averaged velocity is therefore 1 m/sec. It can be shown that the maximum instantaneous velocity is approximately 1.65 m/sec for a 35mm crank length. The relative velocity is a crucial factor in determining the viability of the concept. The piston rings - 31a, 31b, 32a, 32b - are in dry rubbing contact with the respective cylinders and will experience wear at a rate which depends on the rubbing velocity and the normal force. It is necessary that the piston rings form an adequate seal even after 1 - 2 years of continuous operation (during which the seals will have travelled between 31.5 km and 63 km) .

While research is likely to yield improved material combinations, a well-proven combination in vacuum pump technology is to have one component consisting of a material such as PTFE (usually containing an additive such as nickel) and the other consisting of anodised aluminium. It has been found empirically that the relative velocity of the components must not greatly exceed 1 m/sec if excessive wear is to be avoided. Wear rates are characterised in terms of Pv, where P is the pressure applied normal to the sliding surfaces and v is the sliding velocity. Pv numbers have to be small to achieve low wear rates - at around I /sec, a normal pressure of the order of 0.001 to 0.01 psi (7xl0^~5 to 7xl0^~4 kgf/cm²) is recommended by DuPont for granular Teflon^® rubbing on steels. This normal force is so tiny that it cannot be set mechanically. In pump mechanisms such as cryogenic coolers or scroll pumps, PTFE-based seals are allowed to wear to a clearance or are fitted into a loose slot which allows the seal to self-align and "float" in marginal contact with the moving surface .

Description of RotoSync Multi-Stage Vacuum Pump

As with OSRV pumps, best low-pressure vacuum performance is obtained when two pumping stages are deployed in series . It is a particular strength of the RotoSync concept that it permits 2 or more stages in the same rotor, with each stage containing a piston assembly which reciprocates as described above. A preferred arrangement for a two stage pump is shown in Figs. 4A-4D in which the same cross sectional view is presented four times, with hatching used to denote the major components. In Fig. 4A, the hatching denotes the pump stator body 30. In Fig. 4C, the hatching defines the rotor 20 which carries two through holes 20a, 20b which each define the cylinders 20m, 20n for each of two pistons 31a, 31b. The through holes 20a, 20b are separated by an intermediate wall 20e having a through passage 20f around the axis 8 of the stator body 30. The second end 20d of the rotor 20 has a through passage 20g around the axis 8 of the stator body 30. In Fig. 4B, the hatching denotes the two pistons 31a, 31b with their ends shaped to match the curvature of the rotor 20 and the curvature of the corresponding stator body 30. Finally, Fig. 4D denotes the main bearings, the crankshaft 27 which controls the position of the pistons, and the gudgeon pins 40, 40a and connecting bodies 43, 43a which couple the piston assemblies 31, 32 to the crankshaft 27.

Note that all dynamic and static seals have been omitted from Fig. 4 for clarity. At least one shaft seal and a multiplicity of static O-ring seals are required to ensure that the machine is hermetically sealed. The drive end 20c of the rotor 20 is supported by a front bearing 24 which is mounted within the stator body 30. The second end 20d of the rotor 20 is supported by a rear bearing 26. The rear bearing 26 is mounted on the proximal section 27a of the stationary crankshaft 27 which is firmly fixed to the stator body 30 by means of a pin 28. The crankshaft 27 passes through the rotor 20 along its centreline and is supported at its distal section 27b by a distal bearing 25 which is mounted within a cavity 20h at the drive end 20c of the rotor 20. The rotor 20 is drilled radially to give two cylinder through holes 20a, 20b to take the two pistons assemblies 31 and 32. The mechanism used to position the pistons is described below. The stator body 30 is pierced through to provide an inlet port 30d and an outlet port 30c of the multistage pump. It is foreseen that the outlet stage will generally be closest to the front bearing 24 for reasons of tolerancing and thermal expansion of the rotor 20. The stator body 30 also contains a duct 23 which acts as the connection between the inlet port 30b of the outlet (or exhaust) stage and the outlet port 30e of the inlet stage. The inlet port 30b facing a first adjacent piston assembly 31 is connected to the outlet port 30e facing the second adjacent piston assembly 32.

In order for the crankshaft 27 to pass through the pistons assemblies 31 and 32, the body of each piston assembly is pierced through by through passages 33 and 34 such as front slots of approximately rectangular shape, as illustrated in Fig. 1, Fig. 2 and Fig. 5. The width of the through passages 33 must be greater than the diameter of the crankshaft 27 to avoid mechanical rubbing, while the height must be greater than {2h + shaft diameter} but these are otherwise under the designer's control. To allow the piston drive components to be fitted, both pistons assemblies 31, 32 must likewise be pierced through with transverse slots 35, 36 at right angles to the front slots 33, 34.

The drive mechanism for each of the two pistons assemblies is identical and is illustrated for piston assembly 31 in Fig. 5. The crank 46 acts as the carrier for a deep groove ball bearing 44. The crank 46 is a slide fit onto the crankshaft 27 (Fig. 2A) possibly via the solid bush 45 which is offset from the centreline of the crank 46 by the amount h. The crankshaft carries a keyway and a key 48 (see Fig. 2B) for each piston assembly - each key engages into its associated crank or bush to prevent rotation of the crank 46 on the crankshaft 27. The position of each keyway defines the timing of the associated piston assembly with respect to the ports on the stator body 30. In the preferred arrangement, the keys are on opposite sides of the crankshaft such that when one piston assembly 31 is at top dead centre (TDC) the other piston assembly 32 is at bottom dead centre (BDC) . In that arrangement, the diametral through holes 20a, 20b are parallel to each other, and two adjacent piston assemblies 31, 32 are connected to the stationary crankshaft 27 in opposed phase, so that they reciprocate in the respective through holes 20a, 20b in opposite directions.

The crank 46 fits inside the inner shell of the bearing 44. The outer shell of the bearing 44 carries a connecting body 43 which is drilled such that a sleeve bearing 41 can be pressed into a hole 42.

An internal slot is machined into one crown of the piston assembly 31 ; the width of the slot is slightly greater than the length of the sleeve bearing 41, such that the connecting body 43 can rock in the transverse slot 35 through the required angle as the piston assembly reciprocates. A through hole 47 takes a gudgeon pin 40 to lock the connecting body 43 and the crank 46 in place.

The crank assembly (crank 46, bearing 44, connecting body 43) is pre-assembled and slipped into the transverse slot 35 until the sleeve bearing 41 lines up between the holes 47, through which the gudgeon pin 40 is inserted to retain the connecting body at the small-end bearing 41. When assembled, the crank assembly 31 is fully contained within the body of the piston assembly, along its centreline, between the pistons 31a, 31b.

To assemble the rotor, the pistons assemblies 31, 32 (each containing the pre-assembled crank assemblies) are slid into the rotor 20, then the drive crankshaft 27 is slid through both cranks 46, 46a until the keys have engaged both cranks and the drive crankshaft 27 is fully engaged into distal bearing 25. The stator body 30 is then built up around the rotor 20 to result in the completed mechanism. In operation, the rotation of the rotor 20 causes the two pistons assemblies 31, 32 to act upon their associated connecting bodies 43, which rotate passively about the stationary cranks 46, 46a and cause the piston assemblies 31, 32 to move synchronously to the appropriate position in the bores 20a, 20b. Each piston assembly is self-valving between the inlet 30b, 30d and outlet 30c, 30e of each stage and makes two compression strokes per revolution. Fig. 4 also shows that the transfer of gas from the inlet stage piston assembly 32 to the outlet stage piston assembly 31 is effected through duct 23 as a simple transfer with no change in gas volume, as the piston assemblies move in anti-phase.

Because the piston assemblies move in anti-phase and are identical in terms of mass distribution, there are no net out-of- balance forces in the centre plane normal to the crankshaft centre- line 8. However, in the plane of the crankshaft through the centreline 8, the motion of the piston assemblies creates an out-of- balance moment about the shaft centre at a frequency equal to twice the rotational frequency. This can be seen in Fig. 3, where the piston moves from a position of maximum out-of-balance at 0° (and 180°) to a position of balance at approximately 90° (and 270°) twice per revolution. The centre of mass of each piston assembly oscillates with a peak to peak amplitude h from the balance position at the centreline to the extreme and the whole motion is within a single hemisphere - the second piston assembly moves in antiphase in the opposite hemisphere. The out-of-balance force is a couple acting at the centre of the crankshaft and cannot be compensated with balance weights in any simple way.

The volumes within the centre of the mechanism can be vented either to the interstage pressure or to the inlet pressure, depending on design requirements.

Leakage paths

If the pistons are polymer-coated or are fitted with piston seals, it is safe to assume that there is no important leak path along the length of the piston body. Fig. 6A illustrates the exhaust stage only of a multistage pump with the stage inlet 30b and outlet 30c positioned as shown. Fig. 6B shows the same stage in cross-section. In a multistage pump, it is the pumping performance of the exhaust stage which determines the overall performance of the pump. As the exhaust stage works at the highest discharge pressure (atmospheric pressure) , it is internal leaks in this stage which dominate everything else.

Referring to Fig. 6, the most significant leak paths are: 1. The leak paths around the periphery of the rotor 20 from the outlet port 30c (assumed to be at atmospheric pressure) to the inlet 30b - these leaks are illustrated by the arrows PL on

Fig. 6B. There are two leak paths, one in the clockwise sense from X to Y and the other in the counter-clockwise sense. These two leaks act in parallel and can be modelled as rectangular ducts. The length of each duct is approximately

{one half of the circumference of the rotor minus the piston diameter}, the width of each duct is approximately equal to the diameter of the piston, and the height of each duct is equal to the mechanical clearance between rotor 20 and stator body 30.

2. The leak path from the outlet port across the end face 20k of the rotor 20 to the inlet 30b of the stage. This leak is illustrated in Fig. 6A by the arrows EL. 3. The leak path from the outlet port of the exhaust stage to the adjacent inlet port of the inlet stage (not illustrated) along the surface of the rotor in the axial direction. This leak is illustrated by the arrow IL (Fig. 6B) and it can also be modelled as a rectangular duct. This leak path is the most important in determining the achievable ultimate pressure and compression ratio. Each of these is treated below.

1. Peripheral leaks

The clearance between the rotor 20 and stator body 30 is determined entirely by manufacturing tolerances. Assembly workers usually require a radial clearance of 75μ to 125μ diametral clearance of 150μ to 250μ when assembling components consisting of concentric cylinders, as is the case for this machine. The coefficient of linear expansion for aluminium alloys is approximately 2.2xl0^"s /°C. For a 140mm diameter rotor, this represents a clearance effect of 1.5μ per degree of differential temperature between rotor and stator. Assuming a 20 °C differential between rotor and stator, this is equivalent to 30μ of lost radial clearance as the rotor heats up differentially. For this machine, it is assumed that the design radial clearance between rotor and stator should be 75μ when the machine is cold. The leak path geometry can be approximated as a wide, long rectangular duct. The problem is to estimate the effective length of the duct, given the nature of the mechanism. Assuming the effective length is about 100 mm, the ultimate inlet pressure of the stage should be in the region of 35 - 40 mbar, giving a final stage compression ratio of K_m=25 - 29. This would result in a 2 stage machine giving an ultimate pressure in the region of 0.1 mbar if all other leaks could be controlled. As in all dry pump mechanisms, clearance control of the final stage is critical and the use of a sacrificial polymer film on the rotor or stator inner surface would give the designer the option of reducing the clearance in this stage.

End-face Leak

The origin of the end-face leak is illustrated in Fig. 6A. The leak is from X to Y across the end face 20k of the rotor 20 and its associated stator body 30. The geometry of the leak is that of a cylinder of diameter equal to the rotor 20 and a height equal to the clearance between the rotor and stator end faces. Because the leak is fed by gas from one edge of the cylinder at X, the geometry of the leak is non- trivial and difficult to model accurately. However the leak rate is expected to be substantial at standard engineering clearances and should be controlled. A new design for a seal is described below in. However, the magnitude of the leak can also be addressed by tolerancing. The obvious choice of reference plane for the assembly of this mechanism is the plane of the stator end face in the exhaust stage; pump assembly should start here with the insertion of the rotor 20 into the stator end face, through a main front bearing 24 and a shaft seal (not illustrated) . This gives the best chance of controlling the clearance between the end face of the rotor and the stator. The designer must ensure that axial thermal expansion of the rotor is towards the high vacuum end of the mechanism, as rotor/stator interference must be avoided but the clearance must be minimised.

It is important to note that the end-face leak need only be minimised such that the unavoidable peripheral leaks are always dominant. A sacrificial polymer coating on the end face of the rotor (or on the stator end face) would allow the designer to reduce the effect of certain tolerances (especially out-of- squareness) , allowing the rotor clearance to be reduced. Clearly, a hard interference with a significant area of the coating must be avoided as the rotor would seize due to friction.

3. Interstage Leak

The leak path from the outlet port 30c to the inlet port 30d across the top of the rotor 20 is potentially the most significant in determining the ultimate performance of the pump because the leak injects a gas load directly into the inlet stage and hence limits the ultimate pressure. Extremely small clearances are required to limit the leak, as full atmospheric pressure is developed across the short sealing distance between the piston bores. A clearance of as little as lOμ is expected to result in an ultimate pressure of about 1 mbar for air. It is very unlikely that such a small clearance can be held in the assembly process and thermal expansion will anyway create the risk of mechanical seizure. A new sealing concept is described below which will be suitable to control the leak. A New Sealing Concept to Control the End-face Leak

The origin of the end-face leak for the new RotoSync dry pump is discussed above. A new concept for controlling the end-face leak is described below and provides a low cost, self-aligning barrier in the path of the leak which greatly reduces its conductance under normal running conditions . There is provided end face sealing between the end faces

20k, 201 of rotor 20 and corresponding end faces of the stator body 30.

Fig. 7A shows an illustration of the stator end-face 60 of the RotoSync previously described. One of its main functions is to carry the main bearing 24.

Into the plane face which is adjacent to the rotor end- face 20k (Fig. 1) , a circular slot 64 is machined covering approximately 340° of arc, with an outer radius just less than the radius r of the rotor. The slot is left uncompleted in the area adjacent to the inlet of the stage which is at interstage pressure. Into the slot 64 is inserted a polymer strip 62 of rectangular cross-section which makes a light interference with the walls of the slot. The insertion depth is unimportant provided that the strip 62 is everywhere protruding from the slot 64 by more than its final amount, with clearance 62a (Fig. 7B) remaining under the strip 62. The overall length of the strip 62 should be slightly less than the overall length of the slot 64, such that there is a clear channel to allow the small volume clearance 62a underneath the strip 62 to be pumped to its operating pressure. Fig. 7B shows a cross-section of one part of the stator end face 60 with the slot and strip 62 illustrated. During the assembly process of the pump, the rotor 20 or a special assembly tool is brought into contact with the strip 62 and pressed into the stator end-face 60 to an accurately known depth. The action of the pressure is to force the sealing strip 62 deeper into its slot 64 and to everywhere align the strip 62 with the surface of the rotor 20 (Fig. 7C) . During final assembly, the rotor is set to the press insertion depth minus a small increment, as shown in Fig. 7D, such that there is a very small clearance 60b between the strip 62 and the rotor 20.

If it is assumed that the normal running clearance between rotor and stator is 75μ, the sealing strip 62 will project above the surface of the stator 30 by a similar amount. The rotor setting procedure will arrange for a cold clearance of the order of 12μ - 25μ, depending on the method used by the designer. As the machine heats up in normal operation, the rotor 20 will expand with respect to the stator 30 and the clearance will reduce. The reduction is minimised by ensuring that the rotor expands towards the high vacuum end of the machine. If the clearance between the rotor and the sealing strip reduces to the point where there is some light rubbing contact, the polymer will wear locally to a minimum clearance.

The overall effect is to reduce the conductance of the end-face leak. Note that it is only important that the seal is effective over the top half of the rotor 20 where the pressure is very high. Over the bottom half of the rotor 20, the pressure is substantially lower and ultimately becomes equal to the interstage pressure which is the pressure of the interstage duct 23.

This type of seal can be used to advantage to seal other mechanisms where end-face leakage is important.

A New Sealing Concept to Control the Interstage Leak

The interstage leak is described above. The control of this leak is crucial in achieving the lowest possible pressures in the RotoSync dry pump mechanism. A new seal concept is described here which permits the interstage leak to be reduced considerably. The concept is illustrated in Figs. 8A and 8B and the position of the complete seal assembly 29 is shown in Fig. 4A.

The new seal takes the form of two circular but incomplete polymer rings 55 and 56, of rectangular cross-section. These rings are manufactured such that their internal diameter is equal to, or slightly smaller than the diameter of the rotor 20. The rings are inserted into a cavity 29a in the stator wall 30, formed by the final assembly of stator shell 50 and stator shell 51, such that the rings run in contact with the peripheral surface 20j of the rotor 20 in the region 52 between the piston bores. The rings 55, 56 are separated from each other by a solid ring 57 which also acts as a support for each polymer ring 55, 56, with a space 58 between the solid ring 57 and the rotor 20. It is important that the resulting fit of the two polymer rings 55, 56 into the cavity 29a is such that both rings can self-align with the position of the rotor 20 in the radial direction but that there is light interference between the rings 55, 56 and the walls of the cavity 29a. The ends of the rings 55, 56 fit against an anti-rotation feature in the stator cavity in the vicinity of the interstage duct (not illustrated) . Note that it is particularly important that the fit of the rings 55, 56 to the rotor peripheral surface 20j is good over the top half of the rotor 20 where the pressure is at its greatest. The importance of good fit is less over the bottom half of the rotor 20.

In operation, the space 58 between the solid ring 57 and the rotor 20 is held at interstage pressure by means of the location of the cut-out sections of the rings which communicate with the interstage duct 23 via a channel 23a illustrated in Fig. 4A. The space 55a, 56a behind each ring is similarly held at the interstage pressure by the same means. Assuming now that the clearance 53 is at atmospheric pressure by virtue of the fact that it is adjacent to the outlet port, a significant flow of gas will flow under the ring 55 depending on its very small but finite clearance to the rotor 20. However, the space 58 is pumped to the interstage pressure and hence the leakage is pumped back into the inlet of the exhaust stage. Computation shows that the largest conceivable leakage under ring 55 does no more than perturb the interstage pressure because the leak rate is small compared to the pumping speed of the stage.

The leakage under the second ring 56 is now driven by the interstage pressure which is expected to be less than 50 mbar. This means that the mass flow under the ring 56 into the pump inlet is very small and does no more than perturb the inlet pressure, again because the leak rate is small compared with the pumping speed of the inlet stage.

Because the rings make a light interference with the stator 20, the rotor 30 will eventually wear the rings to a marginal clearance and wear will cease. The ability to self-align during the assembly process is critical to ensure that initial rubbing forces are light.

Claims

1 - Dry vacuum pump, having two or more stages in a single rotor (20) , each stage having a solid double ended piston assembly (31, 32) reciprocating in respective diametral through hole (20a, 20b) of the rotor (20) and constrained to move synchronously within the through hole (20a, 20b) which is itself in rotation with the rotor (20) and in line with inlet port (30b, 30d) and outlet port (30c, 30e) of a surrounding stator body (30) . 2 - Dry vacuum pump according to claim 1, comprising :

- a stator body (30) with at least one cylindrical inside cavity (30a) having an axis (8) and opposed inlet ports (30b, 30d) and outlet ports (30c, 30e) ,

- a rotor (20), rotating within the inside cavity (30a) of the stator body (30) around the axis (8) of the inside cavity

(30a), and having at least two diametral through holes (20a, 20b), each diametral through hole (20a) having two opposed large diameter bores which form two piston cylinders (20m, 20n) , the rotor (20) having a drive end (20c) and an opposed second end (20d) , the diametral through holes (20a, 20b) being separated by an intermediate wall (20e) having a through passage (20f) around the axis (8) of the stator body (30), the second end (20d) having a through passage (20g) around the axis (8) of the stator body (30),

- two solid piston assemblies (31, 32) , each of them comprising two opposed pistons (31a, 31b ; 32a, 32b) , slidingly mounted in the respective piston cylinders (20m, 20n) of the respective diametral through hole (20a, 20b) of the rotor (20) , a stationary crankshaft (27) , having a proximal end (27a) which is fixed rigidly in the stator body (30) , and having at least two cranks (46, 46a) which are laterally offset from the axis

(8) of the inside cavity of the stator body (30) by an amount h and which are operatively connected to the respective piston assemblies

(31, 32) so that the piston assemblies (31, 32) reciprocate synchronously along the respective diametral through holes (20a, 20b) of the rotor (20) when said rotor (20) is rotated in the stator body (30) and around the crankshaft (27), the piston assemblies (31, 32) have corresponding transverse through passages (33, 34) between the pistons thereof, the crankshaft (27) crosses through said through passages (20f, 20g), of the rotor (20) and said through passages (33, 34) of the piston assemblies (31, 32), and the cranks (46,

46a) are inside the corresponding through passages (33, 34) of the piston assemblies (31, 32) .

3 - Dry vacuum pump according to claim 2, wherein each crank (46) is operatively connected to the respective piston assembly (31) by means of a connecting body (43) having a first end

(43a) and a second end (43b) , with the first end (43a) that is pivotably secured to the piston assembly (31) by a transverse gudgeon pin (40), and with the second end (43b) that can freely rotate around the crank (46). 4 - Dry vacuum pump according to claim 3 wherein the second end (43b) of the connecting body (43) is joined to the crank

(46) by means of a bearing (44) .

5 - Dry vacuum pump according to claims 2-4 wherein the crank (46) is slide fitted on the crankshaft (27) but prevented from rotating by means of a key (48) .

6 - Dry vacuum pump according to claims 2-5 wherein the drive end (20c) of the rotor (20) is supported by a front bearing

(24) on the stator body (30), the opposed second end (20d) is supported by a rear bearing (26) on the proximal section (27a) of the stationary crankshaft (27) .

7 - Dry vacuum pump according to claim 6 wherein the stationary crankshaft (27) has a distal section (27b) which is supported by a distal bearing (25) that is mounted within a cavity (20h) of the drive end (20c) of the rotor (20) . 8 - Dry vacuum pump according to claims 2-7 wherein the diametral through holes (20a, 20b) are parallel to each other, and two adjacent piston assemblies (31, 32) are connected to the stationary crankshaft (27) in opposed phase, so that they reciprocate in the respective through holes (20a, 20b) in opposite directions.

9 - Dry vacuum pump according to claim 8 wherein the inlet port (30b) facing a first adjacent piston assembly (31) is connected to the outlet port (30e) facing the second adjacent piston assembly (32) .

10 - Dry vacuum pump according to claims 2-9 comprising end face sealing between the end faces (20k, 201) of rotor (20) and corresponding end faces of the stator body (30) .

11 - Dry vacuum pump according to claim 10 wherein the end face (20k, 201) of the stator body (30) has a circular slot (64) with outer radius just less than the radius r of the rotor (20) , a polymer strip (62) of rectangular cross section is inserted into the slot (64), with a clearance (62a) remaining under the strip (62), and with a very small clearance (60b) between the rotor (20) and the strip (62) .

12 - Dry vacuum pump according to claims 2-11 comprising an interstage seal assembly (29) between the outer peripheral face (20j) of the rotor (20) and the stator body (30) in the region between the diametral through holes (20a, 20b) of the rotor (20) .

13 - Dry vacuum pump according to claim 12 wherein the interstage seal assembly (29) comprises two circular but incomplete polymer rings (55, 56) of rectangular cross section, inserted into a cavity in the stator wall, the rings (55, 56) being separated from each other by a solid ring (57) with a space (58) between the solid ring (57) and the rotor (20) .

14 - Dry vacuum pump according to claim 13 wherein the space (58) is held at interstage pressure by means of communication (23a) with the interstage duct (23) .