2 3 This invention relates to rotary seals and bearings, and is a development of the

5 technology described in WO-95/35457 and in WO-97/13084.

6

7

8 BACKGROUND TO THE INVENTION

9 0 Described in those patent publications are structures that serve as rotary seals and I bearings, especially for impeller-type pumps. The structures are characterised as 2 including a pair of sleeves, being a rotor-sleeve and a stator-sleeve, the sleeves having 3 respective interface surfaces; the rotor and stator interface surfaces are disposed in j male/female configuration; and the interface surfaces are so arranged as to define a 5 hydrodynamic-bearing-interface therebetween, during operation of the apparatus, and 6 upon supply of a liquid to the interface. The interface is of a progressively-reducing-

17 diameter (or tapered) configuration.

18

19 One of the interface surfaces is provided with a groove, which extends spirally around

20 and along the axial length of the interface. The groove has an entry-mouth and an exit-

21 mouth. The apparatus includes an entry-chamber, which is in liquid-flow-communication

22 with the entry-mouth of the groove, and the apparatus includes a supply of a liquid in the

23 entry-chamber. The apparatus includes an exit-chamber, which is in liquid-flow-

24 communication with the exit-mouth of the groove, for receiving liquid emanating from the

25 exit-mouth.

26

27 The disposition of the groove in the apparatus is such that, during operation, liquid from

28 the entry-chamber is urged by rotation of the rotor into the entry-mouth of the groove,

29 and is driven along the groove towards the exit-mouth, and into the exit-chamber.

30

31 As shown in the aforesaid prior publications, two of the pairs of tapered sleeves may be

32 provided, arranged in an end-to-end configuration. In that case, thrust forces on the

33 shaft can be contained between the two sleeve-pairs. Thus, the end-to-end

configuration of sleeves can then serve as the whole bearing support required by the shaft.

It is preferred that the liquid or barrier-liquid provided for the hydrodynamic film at the sleeves interfaces is water, or water-based. (In fact, a 50/50 mixture of water with glycol has shown excellent performance). Water, even when mixed with glycol, is not an ideal lubricant; that is to say, it does not form hydrodynamic films so readily as, say, lubricating oil. Also, a hydrodynamic film formed in water is fragile, compared with a film formed in oil; that is to say, it is all too easy to overload the surfaces defining the film, causing the film to break down, whereby the surfaces of the rotor and stator sleeves can touch, metal-to-metal.

The invention as described herein is aimed at providing a manner of arranging and profiling the sleeves, and the interactive surfaces thereof, whereby the possibility of a metal-to-metal contact occurrence is rendered less likely.

GENERALFEATURESOFTHE INVENTION

At each location along the length of the interface, the interface has a respective cone angle, and, in respect of axially-spaced locations A and B along the axial length of the interface, the invention lies in providing that the cone angle of the interface at location A is substantially steeper than the cone angle at location B.

It may also be stated that, at each location along the length of the interface, the interface has a respective film-thickness, being the thickness of the hydrodynamic film in the interface at that location, and, in respect of axially-spaced locations C and D along the axial length of the interface, the thickness of the film at location C is substantially thicker than the thickness of the film at location D.

Preferably, the way in which this change in the steepness of the cone angle, and the change in the thickness of the film, is achieved is curving the profile of the rotor-sleeve / stator-sleeve interface.

THE PRIOR ART

It is well-known to provide grooves in the interface surfaces of hydrodynamic-film bearings. Also, curved bearing surfaces are known, especially spherical bearing surfaces. Examples are shown in US-1 ,132,759 (Bache, 1915) and US-4,614,445 (Gerkema, 1986). Also of interest is the Paper Analysis and Design of Spiral-Groove Bearings by E A Muijderman, in the July 1967 Journal of Lubricating Technology, Transactions of ASME.

Grooves in bearing surfaces are normally provided for purpose of equalising out a hydrostatic pressure in a lubricant. The lubricant is being forced between the bearing surfaces, under pressure, and the grooves ensure the lubricant is spread evenly over the whole surface. Generally, in a grooved hydrostatic /hydrodynamic bearing, the pressure is at a maximum at the place where the lubricant enters the interface, and the pressure gradually drops off as the lubricant moves away from that location.

THE INVENTION IN RELATION TO THE PRIOR ART

In the aforesaid WO-95/35457 and WO-97/13084, barrier-liquid is driven axially along the groove, by the rotation of the rotor, from near zero (gauge) pressure at the entry- mouth of the groove to a maximum pressure (in some cases, as high as 100 psi and more) at the exit-mouth of the groove. The pressure of the barrier-liquid increases progressively (and, in some cases, substantially linearly) as the barrier-liquid passes through the bearing.

If the event of a breakdown of a fragile water film, it may be surmised that the failure of the film will occur first at the location where the hydraulic pressure in the film is lowest. Therefore, the designer preferably should seek to arrange, if possible, that in the portion of the interface where the pressure is lowest, the sleeve surfaces are a little further apart - or in other words, that the film is a little thicker in that portion.

In a case where the hydrodynamic-water-film interface between a pair of sleeves is supporting a heavy thrust force, if the cone-angle of the interface is steep in one location and not-so-steep in another location, it may be surmised that the thrust forces are being supported more in the location where the cone is the steepest. Thus, again in an effort to defer the onset of metal-to-metal contact, the designer should seek to provide, at the locations where the cone angle is steepest, that that is a location where the surfaces are a little further apart.

As will be described herein, by making the profile of the interface surfaces of variable or changing cone-angle, the properties of the interface from the standpoint of thrust- support, and the thickness of the film, can be tailored by the designer to suit the circumstances.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS

By way of further explanation of the invention, exemplary embodiments of the invention will now be described with reference to the accompanying drawings, in which:

Fig 1 is a cross-sectioned side elevation of a rotary impeller pump, which embodies the invention; Fig 2 is a similar view of another pump that embodies the invention; Fig 3a is a close up of a pair of sleeves from the pump of Fig 1 ; Fig 3b is the same view as Fig 3a, but shows the sleeves in a separated condition; Fig 4a is a close up of a pair of sleeves from the pump of Fig 2; Fig 4b is the same view as Fig 4a, but shows the sleeves in a separated condition; Fig 5 is a cross-sectioned side elevation of a rotary machine, which embodies the invention; Fig 6a is a close up of a pair of sleeves from the machine of Fig 5; Fig 6b is the same view as Fig 6a, but shows the sleeves in a separated condition; Fig 7 is a cross-sectioned side elevation of another rotary impeller pump, which embodies the invention; Fig 7a is a similar view of another rotary impeller pump, which embodies the invention;

Fig 8 is a similar view of another rotary impeller pump, which embodies the invention; Fig 8a is a similar view of another rotary impeller pump, which embodies the invention; Fig 9 is a similar view of another rotary impeller pump, which embodies the invention; Fig 9a is a similar view of another rotary impeller pump, which embodies the invention; Fig 10a is a diagrammatic, simplified cross-sectioned side elevation of the pump shown in Fig 7; Fig 10b is a similar view of the pump shown in Fig 8; Fig 10c is a similar view of another pump; Fig 10d is a similar view of another pump; Fig 1 1 a is a similar view of another pump; Fig 1 1 b is a similar view of another pump; Fig 11 c is a similar view of another pump; Fig 12a is a similar view of another pump; Fig 12b is a similar view of another pump; Fig 13 is a cross-sectioned side elevation of a reciprocating machine, which includes the invention; Fig 14 is a similar view of another machine that includes the invention.

The apparatuses shown in the accompanying drawings and described below are examples which embody the invention. It should be noted that the scope of the invention is defined by the accompanying claims, and not necessarily by specific features of exemplary embodiments.

The pump 20 as shown in Fig 1 includes a rotor sleeve 23 mounted on a shaft 24, and a stator sleeve 25. The outward-facing surface 26 of the male rotor-sleeve 23 and the inward-facing surface 27 of the female stator-sleeve 25 together define an interface. A spiral groove 28 is provided in the male rotor surface 26, the groove having an entry- mouth 29 and an exit-mouth 30. The entry-mouth 29 communicates with an entry- chamber 32, and the exit-mouth 30 communicates with an exit-chamber 34. It may be noted that the surface 26 contains only the one groove 28, the groove having several turns. The hand of the spiral groove 28 is such that upon rotation of the shaft 24, liquid present in the groove 28 travels along the groove from left to right in Fig 1 .

Barrier-liquid is supplied at ambient pressure to the entry-chamber 32 from a reservoir 35.

As shown in Fig 1 , the exit-chamber 34 is open to the process-chamber 36. In other cases, the exit-chamber can be sealed from the process-chamber, as will be observed in some of the accompanying drawings.

The shaft 24 is provided with bearings shown diagrammatically at 37. The interface between the surfaces 26,27 serves as a seal, for sealing the process fluid from the atmosphere. The female stator-sleeve 25 is spring- (and pressure-) biassed into contact with the male rotor-sleeve 23. Even though the interaction of the sleeves 23,25 is intended as a seal, not as a bearing, of course the sleeves do provide some support for the shaft.

The grooved male rotor surface 26 is curved convexly as shown in Fig 1. In Fig 2, the situation is similar, except that the grooved male rotor sleeve is curved concavely.

In Fig 2, the supply of barrier-liquid to the entry-chamber 32 is not shown, and the entry chamber is sealed from the atmosphere by packing 38.

Fig 3a shows the sleeves 23,25 of Fig 1. The cone angle of the interface at location A is steeper than the cone angle at location B. The cone angle at A is substantially more than ten degrees (included) and the cone angle at B is less than ten degrees.

Fig 3b shows the sleeves somewhat separated. Although Fig 3 exaggerates the effect of axially separating the sleeves, it will be noted that the thickness of the film at axial location C increases more than the thickness of the film at location D, as the sleeves are separated. Figs 4a and 4b show the same thing in respect of Fig 2.

Fig 5 shows a machine with bearings supporting a shaft. Here, two pairs of rotor- and stator-sleeve are shown, in an in-series end-to-end or back-to-back configuration. The sleeves define a spherical interface, which allows the shaft to be in angular- misalignment with respect to the frame of the machine (in the manner of conventional

spherical bearings).

Figs 6a and 6b illustrate the progressive nature of the changes in film thickness as the sleeves of Fig 5 are axially separated.

Fig 7 shows a pump, having an impeller 40 operating in a process-fluid-chamber 42. The impeller is driven by a shaft 43, which takes torque from a motor (not shown) to the left. The shaft 43 receives only torque from the motor, and the shaft has no other bearings, other than the sleeves as shown, which support the shaft against journal (radial) and thrust (axial) forces.

In operation, the rotor sleeves 45L,45R are solid with the shaft 43. The stator sleeves 47L.47R are solid with the housing 49. Mechanical seal 50 seals the process-chamber 42 from the sleeves, and mechanical seal 52 seals the sleeves from the atmosphere.

In Fig 7, the sleeves are arranged in end-to-end configuration. The sleeves are centre- fed, in that barrier-liquid is admitted (substantially not under pressure) into a chamber 53, which serves as the entry chamber for both sleeve-pairs 48L (45L+47L) and 48R (45R+47R) for conveying liquid to the entry-mouths of the grooves. The spiral grooves 54L.54R are oppositely handed, whereby the liquid is forced to the left in the left sleeve- pair 48L, and to the right in the right sleeve-pair 48R, during rotation. From the exit- mouths of the grooves, the liquid emerges, now under pressure, into the respective exit- chambers 56L.56R.

Fig 7a shows a similar arrangement of sleeves, but now the barrier-liquid is end-fed. That is to say, the liquid is supplied to an entry-chamber 57 at the extreme left, passes along the interface of the first sleeve-pair 48L, progressively gaining pressure, and emerges into the intermediate chamber 58. From there, the liquid passes along the interface of the second sleeve-pair 48R, gaining more pressure, and emerges into the exit-chamber 59. The intermediate chamber 58 may be regarded as the exit-chamber for the first sleeve-pair 48L and as the entry-chamber for the second sleeve-pair 48R.

The behaviour of the apparatuses with the two ways of feeding the barrier liquid is

different, as will now be explained.

When the pump is running, and process fluid is passing through the impeller, the momentum of the flow gives rise to a reaction force, which draws the impeller 60 to the right. This force appears as an axial force on the shaft, and is supported by the shaft bearings. The sleeve-pairs 48L.48R serve as the shaft bearings in Figs 7 and 7a, and therefore, during normal pumping, the sleeve pair 48R is under a constant thrust force, from the impeller-draw. In Figs 7 and 7a, the sleeve-pair 48L sees no thrust force from impeller-draw.

If the supply of process fluid in the intake to the impeller should fail, or partially fail (e.g under conditions of cavitation), the thrust force on the shaft disappears or is reduced, with consequent effect on the sleeves.

In this specification, the term cavitation refers to the condition where the impeller is rotating but there is an interruption in the supply of process fluid in the intake to the impeller. Cavitation can be partial or (rarely) complete. When cavitation is partial, the thrust force due to impeller-draw is reduced: however, in practice, cavitation can cause vibrations, and spurious effects, whereby the thrust force might even be (momentarily) increased. If cavitation is complete, the impeller-induced thrust-force on the shaft theoretically drops to zero.

Also, in addition to the impeller-induced thrust-forces acting on the shaft, it should be borne in mind that the pressures in the various chambers also acts on the sleeves, urging them in the appropriate directions. In Fig 7 (centre-fed), when there is no thrust on the shaft (i.e under complete cavitation) the pressure in the two exit chambers 56L.56R can be expected to be the same, whereby the centre-fed shaft (theoretically, anyway) sits neutrally during complete cavitation.

But in Fig 7a, the pressure in the intermediate chamber 58 is intermediate, and the pressure is high in the exit-chamber 59. The pressure in the chambers gives rise to a force tending to drive the two sleeves, and with them the shaft, to the left. Therefore, when the pump is (completely) cavitating, and there is little or no impeller-induced draw

on the shaft, the end-fed shaft is urged strongly to the left by the pressures in the chambers.

Thus, when the pump is pumping process fluid normally, as noted, the impeller draws the shaft to the right. In Fig 7a, the fact that the pressure in the chambers is also urging the shaft to the left can serve as a counter to the impeller-induced thrust force on the shaft. In an ideal case, in an end-fed mode (Fig 7a), the rightwards thrust force due to impeller-draw can be exactly balanced by the leftwards thrust force due to the pressure in the chambers.

It may be considered that in Fig 7, the centre-fed case, the shaft is (nearly) neutrally- thrust-balanced under complete-cavitation conditions, i.e when there is no process fluid in the intake to the impeller, but the shaft is urged to the right by impeller-draw under normal pumping; whereas in Fig 7a, the end-fed case, the shaft is (nearly) neutrally- thrust-balanced under normal pumping, but suffers a thrust force to the left, because of the pressure in the chambers, under cavitation.

The designer can choose to feed the barrier-liquid into the sleeves according to which mode of operation best suits the particular installation.

As will be understood by perusing Figs 6a,6b, in the centre-fed case (Fig 7), the film in the interface of the right sleeve-pair 48R, which under normal pumping is the sleeve-pair that is reacting the thrust force due to impeller-draw, is thicker near the impeller, where the cone angle of the interface is steeper, and the film is thinner at the other end of the right sleeve-pair 48R, where the cone angle is less steep.

The proposition can be put forward that, at each axial location along the length of the interface, the incremental increase in pressure per incremental change in axial location is inversely proportional to the film thickness at the location. That is to say, the pressure increases in the shortest distance where the film is thinnest. The thicker the film, the less the pressure increases with length; the thinner the film, the more the pressure increases with length.

So, for an interface that is to take thrust, the designer should see to it that the film is at the higher pressure at the location where the film is predominantly taking the thrust, and also that the film is thickest at the location where the film is predominantly taking the thrust.

In the sleeve-pairs as described, the pressure of the barrier liquid increases progressively from entry end to exit end. Therefore, for a sleeve-pair that is taking thrust, the designer should follow this first design rule: put the steep slope at the exit end, and put the thick film at the exit end.

Consider how this rule applies in Fig 7. During normal running, the right sleeve-pair 48R is taking the thrust due to impeller-draw. The exit end (i.e the right-hand end, of the right sleeve-pair is more steeply sloping, and the right-hand-end is where the film is thickest. So the mode of curvature of the right sleeve-pair 48R in Fig 7, according to the rule just expressed, is the correct way round, i.e rotor-convex is correct.

In Fig 7, under conditions of complete-cavitation, i.e when there is no process-fluid in the intake to the impeller, neither of the sleeve-pairs takes thrust, and so the shape of the curvature is less important. The left sleeve-pair 48L in Fig 7 nominally does not take thrust under either condition, and its shape is less important.

In Fig 7a, the end-fed configuration, again the right sleeve-pair reacts the thrust due to impeller-draw. The high pressure in the exit-chamber 59 however drives the shaft to the left, thus relieving some of the load from the right sleeve-pair interface. As mentioned, in some cases, the two thrust conditions might exactly balance, whereby neither sleeve- pair is taking thrust, in which case their exact shapes do not matter. But in Fig 7a, under cavitation, the pressure in the chambers drives the shaft to the left, and under this condition it is the left sleeve-pair that is under thrust. Therefore, if the left sleeve-pair is designed according to the first design rule above, the cone angle should be steeper, and the film should be thicker, at the exit end of that sleeve interface.

Now, the left sleeve-pair 48L as shown in Fig 7a does not meet this condition. Here, the exit end of the left sleeve-pair is the intermediate chamber 58. That is where the

pressure in the left sleeve-pair interface is higher, and that is where the film should be thicker, and the slope should be steeper.

For operation in the end-fed mode, the shape of the curvature of the left-sleeve should rather be designed as shown in Fig 8.

Since the right sleeve-pair nominally never does take thrust when end-fed (Fig 7a), it does not matter so much what its shape is. Of course, loads and conditions are not always perfectly predictable, and the arrangement of Fig 8 follows the rule as to steepness of slope and thickness of film, whichever of the sleeves happens to be taking thrust. As mentioned, under cavitation conditions, vibrations and partial loadings can cause unpredictable or short-lived spikes of thrusts; but these are catered for very easily, as shown.

It is not suggested that the above design rule — i.e that a sleeve-pair under thrust should be curved so as to have the thicker film and the steeper slope at the higher pressure end of the sleeve-pair - is applicable to every situation. A second way of looking at the potential modes of failure (i.e of the onset of film-breakdown and metal-to-metal contact) is that contact is more likely to occur at the location where the film is thinner; therefore, the thinner film should be at a place where the pressure is higher, since a thin film at high pressure is less likely to break down than the same thin film at low pressure. Therefore, the second design rule would be to the effect that the thinner film should go at the exit end of the sleeve-pair, where the pressure is higher.

Following this second design rule, for the right sleeve-pair in Fig 7 the shapes shown therein would be contra-indicated, in that in that sleeve-pair the film is thinner at the entry end of the sleeve-pair. According to the second rule, the right sleeve-pair should have been rotor-concave. Fig 8a shows a version that follows the second design rule.

For the left sleeve pair in Fig 7a (end-fed), which comes under increasing thrust load with increasing cavitation of the pump, according to the second rule, the configuration shown in Fig 7a, i.e the rotor-is-convex configuration, would be preferred over that of Fig 8. Again, it does not matter so much what shape the right sleeve-pair has, since that is

not under thrust during cavitation, and is not under so much thrust, even under normal pumping. Besides, in Fig 7a, the entry pressure of the liquid fed to the entry mouth of the groove of the right sleeve-pair has already been elevated in pressure by the action of passing through the left sleeve-pair.

The two design "rules", though of course incompatible with each other, can both have applicability under different conditions. For example, if cavitation is only a remote possibility, the end-fed mode should be favoured by the designer. It can be regarded that the main function of the left sleeve-pair in that case is to pre-pressurise the liquid entering the right sleeve-pair; therefore, the second rule does not apply, and the right sleeve-pair should be designed according to the first rule, i.e it should be curved rotor- convex, as shown in Fig 7a.

On the other hand, cavitation is a common bane of pumps; it is all too usual that if a pump fails, it fails because of what happens during cavitation. In a case where that is so, the centre-fed mode should be favoured. If the cavitation is complete, and there is no process fluid at all in the impeller intake, the shaft being then under neutral thrust, both sleeve-pairs are providing only a minimum of pressure, and the second rule assumes more importance. In that case, the thickest films should be at the entry ends of the sleeve-pairs; therefore, both sleeve-pairs should be rotor-concave (Fig 8a).

It is noted that the provision of the curved profiles of the sleeve interfaces permits a reprieve or deferment from the onset of film breakdown: the designer should be aware that the enhancement can arise sometimes when the curve is rotor-convex, but other times when the curve is rotor-concave.

In the structures shown in Figs 7 and 7a, the normal-pumping thrust force due to impeller-draw is reacted in the right hand sleeve-pair, i.e in the sleeve-pair nearest the impeller. This is preferred, in that the thrust-loaded sleeve-pair has a smaller running clearance than the thrust-unloaded sleeve-pair.

On the other hand, it can be arranged that the left sleeve-pair is the one that takes the thrust force on the shaft due to impeller-draw. In that case, the left and right sleeve-

pairs are arranged end-to-end in "bow-tie" configuration, as shown in Fig 9. (The end- to-end configuration shown in Figs 7,7a,8,8a may be termed "football" configuration, by contrast.) Fig 9a also shows a variety of bow-tie configuration.

Figs 10a, 10b, 10c, 10d show diagrammatically the arranged convexity / concavity of the curvatures of the sleeve-pair interfaces, in following the first design rule, i.e the cone angle should be steepest, and the film should be thickest, at the high-pressure end of the sleeve-pair. Fig 10a shows football / end-fed; Fig 10b shows football / centre-fed; Fig 10c shows bow-tie / end-fed; and Fig 10d shows bow-tie / centre-fed.

It is also recognised that it is possible to achieve the required changes in cone angle steepness and film thickness, not by curving the profile of the interfaces, but by arranging two (or more) straight slopes along the length of the interface. Fig 1 1 a shows such an arrangement, which corresponds to Fig 8. Fig 1 1 b shows an arrangement that corresponds to Fig 7.

In fact, providing two slopes can introduce another design flexibility, in that there is no need for the thicker film to be at the same location as the steeper slope. Fig 1 1 c shows the thicker film at the entry ends of both sleeves, but the steeper slope is at the exit end. This arrangement might be favoured in some cases, especially since it can in this case be arranged that, if the sleeves were to come together in metal-to-metal contact, they would touch first at the exit end, where the pressure is higher.

In the case of the as-shown smoothly-curved sleeve-pairs, either convex or concave, if the sleeves are taken out of the pump and pressed together, the interface surfaces make contact with each other over substantially the whole of the interface simultaneously. That is to say, the film thickness is equal over the whole curved interface only when the film thickness is zero.

However, that is not an essential condition of the invention, as has been demonstrated in Fig 11 c. The same condition as has been provided by the two straights in Fig 1 1 c can be machined into the surfaces even when the surfaces are curved. Thus, even when the surfaces are curved (as opposed to two-straights as in Fig 1 1 c), the

curvatures, i.e the cone-angles, can be set to follow both design rules. The curvature can be set so that the film thickness varies from thickest-at-the-low-pressure-end-when- the-fiim-is-thin, to thicker-at-the-high-pressure-end-when-the-film-is-thick, within the same sleeve-pair.

When the surfaces are so machined that they do not touch all-over at zero-clearance, of course the surfaces cannot be lapped against each other, to give a good fit. But modern computer-controlled manufacturing lathes can turn surfaces very accurately in any event, whereby lapping is generally not needed.

Fig 12a shows a sleeve-pair in which the curvature changes between rotor-convex and rotor-concave, within the one interface. It may be noted that, in the absence of barrier- liquid, the interface surfaces would come together simultaneously all-over, and yet, as shown in Fig 12b, the configuration is very favourable for handling thrust on a sleeve- pair. The film thickness is wide at the entry, or low-pressure, left end region of the interface, so that there is plenty of opportunity for the liquid to establish itself into the entry-end of the interface even when contact is imminent. That is to say, the interface is not starved or cut off from the supply of liquid in the entry chamber. Once established in the interface, the liquid then passes to the central region of the interface, where the clearance is small, and therefore where the pressure rises rapidly as the liquid flows from left to right. Finally, at the exit, or high-pressure, right end region of the interface, the pressure in the film is high, the film is thick, and the cone angle is steep, which, as mentioned, are excellent conditions for supporting heavy thrust forces.

The benefits of curve-profiling the sleeve-pair interfaces are not limited to pump installations. Fig 13 shows a compressor, in which the big-end bearings on the crank- pin are curve-profiled. The spherical configuration shown here is advantageous, because it permits the connecting rod of the machine to exactly follow the line of action of the piston in the cylinder, even if the cylinder might not be quite square to the axis of the crank-pin. The arrows on the drawing indicate the direction of flow of the lubricant through the various zones. It should be recalled that the action of the sleeves generates pressure in the lubricant as the lubricant passes through the sleeve-pair, whereby the lubricant emerges from the interface at a higher pressure than it entered, and so the

need for high hydrostatic pressurisation of the lubricant is eased.

Fig 14 shows curved-profiling of the main-bearings, and a rotor-concave profile for the big-end bearings.