GB2230063A

GB2230063A - Multi-element shaft seal

Info

Publication number: GB2230063A
Application number: GB8907627A
Authority: GB
Inventors: Norman Kenneth Bowers
Original assignee: Vickers PLC
Current assignee: Vinters Ltd
Priority date: 1989-04-05
Filing date: 1989-04-05
Publication date: 1990-10-10
Also published as: GB9119877D0; WO1990012229A1; AU5416290A; GB2250067A; GB2250067B; GB8907627D0

Abstract

A seal assembly for sealing a space between a rotary shaft (44) and a stationary member (10) includes a housing (66) having a bearing (90) that maintains the housing (66) true on the shaft (44). Seal assemblies (103, 105 and 107) axially spaced along the shaft (44) are connected via pressure relief valves (134) (Fig. 2) so that e.g. pressure above a predetermined limit across a first ring (103) passes to a second ring (105) and when the threshold at ring (105) is exceeded, pressure passes to third ring (107). In this way, at high pressures all three seal assemblies (103, 105 and 107) are active. A normally inactive back-up seal (109) may be provided. The seal assembly may be water lubricated. A split lap-jointed structure for a sleeve (110) forming part of the seal assemblies (103 to 109) and the curvature of the sleeve (110) relative to shaft diameter are also described.

Description

MULTI-ELEMENT SHAFT SEAL This invention relates to apparatus for sealing a space between a rotary shaft and a stationary member surrounding the shaft.

A problem with which the invention is concerned is how to seal a rotary shaft against a high external or internal pressure where the shaft has a plurality of sealing rings axially spaced from one another. A shaft seal having axially spaced sealing rings is described in Patent Specification No. EP-B-0125896 (James Walker & Co., Ltd) and is believed to be used primarily for sealing between casings and shafts of pumps and turbines, but there is no provision for sharing the pressure drop across each of the rings, so that the seal with the lowest leakage rate becomes subject to the whole pressure drop.

The plain region of the shaft may be defined by a shaft sleeve on which each sealing ring and the bearing means act. The shaft sleeve may need to be made of a wear- resistant and non corrodable material such as Inconel (an iron-chromium-nickel alloy) coated with a hard layer such as tungsten and/or chromium carbide e.g. Union Carbide LW5 material, steel and may be attached to the shaft by electrically insulating connectors to avoid cathodic attack.

The present invention provides apparatus for sealing the space between a rotary shaft and a stationary member surrounding the shaft, wherein the shaft is sealed by means of a plurality of seal elements and means controls the pressure exerted on each element so that at least at high pressure the pressure drop across the apparatus is distributed between the elements.

Conveniently means is provided for regulating the pressure drop across each seal element.

In an alternative aspect, the invention provides apparatus for sealing the space between a rotary shaft and a stationary member, comprising at least two seals spaced axially apart along the shaft to be exposed to a higher pressure and to a lower pressure, wherein means connected between the high pressure sides of the seals limits the pressure difference across the high pressure seal so that if pressure across the apparatus exceeds said limit, that pressure is shared by the seal on the low pressure side.

The number of seals that may be spaced axially along the shaft and connected in series at their high pressure sides by pressure limiting devices will depend on the expected load. For example, in a pratical form of the apparatus designed to resist pressures of the order of 40 bar, there may be three such seals connected in series via two devices for limiting the pressure difference, the pressure across each seal being up to about 10 to 13 bar. To cater for greater pressure, the number of series-connected seals may be increased as required. Alternatively, more seals may be provided if it is desired to produce the working pressure across each seal to e.g. 5 bar.

The problem with the intermediate or low-pressure side seals is that there may in use be only a minimal flow of pressure fluid between the seal and shaft, such flow being relatively independent of the pressure at which the seal is operated. To maximise heat removal, the intermediate and low pressure seals, and also advantageously the high pressure seal, may be housed in retaining rings of relatively high thermal conductivity such as gunmetal.

As a precaution against seal failure, a seal may be provided on the low pressure side that is not series connected to the preceding seals. In that case, that further seal may have a high pressure side connected to a drain passage so that abnormal flow through that drain passage indicates failure of the preceding seals on the shaft. The seal at the low pressure side may be housed in a retaining ring of lower thermal conductivity but higher mechanical strength than the high thermal conductivity seal retaining rings e.g. of aluminium bronze.

The seal on the low pressure side can also be series connected to the preceding seals and yet still be provided with a connection to a drain passage to indicate failure of the preceding seals. This is acceptable where the set working pressure across the seals other than the outermost seals total more than the maximum working pressure. The drain connection to the outer seal may then be fitted into a shut-off valve so that this seal may be brought into use when required.

Patent Specification No EP-B-0125896 shows a seal for a rotary shaft having a flexible sleeve of elastomeric material that in the relaxed state has a diameter greater than that of the shaft so that a positive clearance exists between the shaft and the sleeve. A problem with a seal of this construction is that it is extremely difficult to maintain the proposed small positive clearance in practice. To do this would mean butting the ends of the flexible sleeve very precisely, making due allowance for the compressibility of the sleeve. Even if this could be achieved initially, the clearance will subsequently be affected by temperature changes and by water absorption into the sleeve material.

It is preferable therefore to have a lapped joint similar to a conventional piston ring joint, so that even in the relaxed state the flexible sleeve is always in contact with the shaft. The lapped joint has a gap (Figures 3 and 4 below) so that the sleeve can accommodate expansion and contraction of the shaft with temperature and also expansion and contraction of the sleeve itself due to temperature changes and water absorption.

In a further aspect the invention provides apparatus for sealing the space between a rotary shaft and a stationary member surrounding the shaft, comprising a housing surrounding the shaft, a sealing ring contained in the housing, and means for applying fluid pressure to the sealing ring thereby to apply radially inward pressure to the sealing ring, wherein the sealing ring has at least one split therein defined by a lapped joint and/or in its relaxed state has a curvature greater than that of the shaft.

Various forms of the invention will now be described, by way of illustration only, with reference to the accompanying drawings in which: Figure 1 is a fragmentary section of a shaft having a seal according to the invention fixed thereto; Figure 2 is a detail showing in section a pressure relief valve fitted to the seal assembly of Figure 1.

Figures 3 and 4 are side views of alternative forms of a lap joint between segments of a sleeve for forming part of the seal of Figures 1 and 2; Figure 5 is a transverse section of the sleeve; and Figure 6 shows diagrammatically a preferred relationship between the curvature of the sleeve and the shaft.

In the drawings the high pressure side of the seal is to the left as viewed in Figure 1, and pressure fluid which it is desired to seal against is generally indicated by the arrow 14. Typically this may be water.

A shaft 44 about 0.7 metre in diameter has a larger diameter region defined by a wear-resistent metallic sleeve 50 which may be of hardened Inconel or hardened Inconel-coated steel. To prevent electrolytic action between the sleeve 50 and the shaft 44, an inner sleeve 46 of electrically insulating material fits between the sleeve 50 and the shaft 44. The sleeve 50 is retained at one end by a split retaining ring 48 of electrically insulating material to which the sleeve 50 is attached by bolts 54. A seal 52 of rubber at the other end of sleeve 50 prevents water ingress under the sleeve 50.

A bearing and seal carrier assembly 60 fits onto the shaft 44. It has a bearing carrier part generally indicated by the reference numeral 62 and a seal ring assembly generally indicated by the reference numeral 64. The bearing carrier part 62 includes a bearing housing 66 which is generally tubular.

A water-lubricated journal bearing 90 which runs on the outer sleeve 50 is defined by a pair of abutting half-sleeves of material which will run on metal with water lubrication such as a rubber material or a phenolic plastics material. The half-sleeves of the bearing 90 fit into a split bearing ring 92 each half of which is attached by bolts 94 to the shaft follower housing 66. The outer surface of the bearing ring 92 is formed with grooves 96 for ease of dismantling.

Water is fed through cooling liquid feed port 98 under pressure to a feed annulus 100 defined by a recess in the outer surface of the bearing ring 92. Distribution passages 102 lead from the feed annulus 100 to a first sealing ring assembly generally indicated by the reference numeral 103. In use, water is pumped by a pump into the feed port 98 at a higher pressure than the pressure of water indicated by arrow 14. Water flows from feed port 98 through annulus 100 and distribution passages 102 to the high pressure side of the first sealing ring assembly 103.

The sealing ring assembly 103 comprises an annular sleeve 110 of a suitable plastics bearing material (e.g. a polyester reinforced phenolic resin) formed in two halves that fit together with non-abbutting lapped ends into an elastomeric ring 104 of U-profile (e.g.

of nitrile rubber) that holds the split annular sleeve 110 onto the hardened Inconel outer sleeve 50 in the manner of a piston ring. A single split ring may be used instead of two half-rings, in which case the ends of the split ring are also lapped but do not abut. The elastomeric ring 104 surrounds the sleeve 110, water pressure applied between the housing 112 and the ring 104 acting to transmit radial pressure to the sleeve 110. The ring 104 is in sealing engagement with an axial face of housing 112 on the low pressure side of ring 104 so that fluid can flow from the high pressure side of housing 112 to a space 121 between the elastomeric ring 104 and an inner wall of the housing that is spaced radially from the wall 104.

Distribution passages 102 open directly into the space 121 and grooves 101 at the high pressure end face of elastomeric ring 104 allow water to flow from the space 121 past the end of bearing ring 92 and to return in the direction of flange 68 to the bearing 90. Water flows from bearing 90 into a cavity 91 defined between the end face of outer sleeve 50 and the adjacent face of flange 68. The possible axial travel of shaft 44 relative to bearing housing 66 is indicated by the axial length 89 of cavity 91.

The seal of the invention may be used to hold back fluid pressures amounting to several tens of atmospheres, typically pressures of 13 bar or above.

For that purpose, several seal assemblies 103, 105 and 107 are arranged at axial spacings along the shaft and at least at high applied pressures are simultaneously energised to divide the pressure difference between them. Means such as pressure relief valves or reducing valves are connected between the high pressure sides of the seals to provide for sharing of the pressure of fluid 14 between them. A back-up seal assembly 109 is preferably provided behind the seal assembly 107. It is normally not energised, but may be energised to act independently if the seal assemblies 103, 105 and 107 should fail.

We have determined that from the standpoints of reliability and seal life, the maximum desirable working pressure across a water-lubricated shaft seal consisting of a sleeve of polyester-reinforced phenolic resin or the like in an elastomeric retaining ring is about 15 bar when the shaft that it is running on is of a poor thermal conductivity material such as Inconel. To provide a shaft seal having a greater working pressure than about 15 bar, it is desirable to provide a number of such seals located adjacent one another and spaced at intervals axially of the shaft, and to provide means such that the seals are simultaneously or sequentially energised and the pressure difference is shared amongst the energised seals.If reducing valves or orifices are used in Figure 1, the seal assemblies 103, 105 and 107 are simultaneously engerised by fluid pressure, seal assembly 103 working from high to intermediate pressure, seal assembly 105 working from intermediate to low pressure and seal assembly 107 working from low positive pressure to ambient pressure and the fluid pressure difference is shared amongst the energised seals. It may be preferred, however, that the pressure across the seals should be regulated by pressure relief valves, in which case the seals become energised sequentially as the pressure across the apparatus is increased, and do so in the inverse order of their leakage rates and not necessarily in the sequence 103, 105, 107. If at the expected working pressure only some seals are energised, the remainder are subject to little or no significant wear.An additional back-up seal assembly 109 is provided which may be brought into use should the seal assemblies 103, 105 and 107 fail.

The first seal assembly 103 is housed in a recess 122 of a first seal retaining ring or housing 112 of thermally conductive material such as gun metal. The recess 122 defines the internal space 121 which is fed with water from distribution passages 102 at a pressure somewhat above the pressure of fluid 14 prevailing on the high pressure side of the assemblies. Seal housing 112 is attached by bolts (not shown) to the bearing housing 66, and the fluid tightness of the space 121 is assured by an intermediate 0-ring seal 120. As is apparent from Figure 2, the first seal housing 112 is formed with flow passages 130, 132 communicating the high and low pressure sides thereof. Flow through passages 130, 132 is controlled by means of a pressure relief valve 134 set by a control spring 136 to open at a differential pressure less than the safe working pressure of the first seal assembly 103 (e.g. a differential pressure of about 13 bar). For convenience of access and adjustment the valve member 134 works in a radial bore 138 of the first seal housing 112, in which the valve member is retained by a bolt 140. Pressure across the seal assembly 103 in excess of 13 bar is communicated via passages 130, 132 to the high pressure side of the second seal assembly 105 which is again retained by an intermediate pressure seal housing 114 of gun metal or other heat conductive material.

The housing 114 is attached to the first housing 112 by means of bolts 115 and contains an elastomeric ring 104 with grooves 101 on its high pressure end face and an annular sleeve 110 of polyester reinforced phenolic resin as described above. In the case of the seal in housing 114, the action of the passages 101 on the high pressure side of the ring 104 is to lead leakage and by-pass water from the first seal assembly 103 to the internal space 121 of the second seal assembly 105. The seal housing 114 likewise has one or more pressure relief valves as shown in Figure 2 leading from its high pressure side 121 to the high pressure side 121 of the third seal assembly 107. Again the effect of a pressure difference greater than about 13 bar across the second sea). assembly 105 is to transmit the excess pressure to the third seal assembly 107.

A third annular seal housing 116 also of gun metal or other heat conductive material and containing a ring 104 and sleeve 110 is bolted to the seal housing 14.

The action of the passages 101 on the high pressure side of ring 104 is again to lead leakage and by-pass water from the second seal assembly 105 to the internal space 121 of the third seal assembly 107. The reason why housings 112, 114 and 116 are preferably made of a conductive material such as gunmetal is in order to remove heat effectively from between the polyester-reinforced phenolic sealing rings 110 and the shaft 50 which is generated by friction on rotation of the shaft 44.

On the low-pressure side of the seal assembly 107 there is provided a back-up seal assembly 109 in a back-up seal housing 118. The seal housing 116 contains a ring 104 and a sleeve 110 but does not contain a relief valve communicating the high pressure sides of seal assemblies 107, 109. Instead, back-up seal housing 118 has a valved leak-off passage 126 that is normally open, but may be closed off in the event of failure of the other seal assemblies 103, 105 and 107 to permit the shaft 44 to continue to run The non-energised sealing ring assembly 109 is subject to negligible wear.

The seal assemblies 103, 105, 107 and 109, like the bearing 120 are lubricated by the flow of water through feed port 98 or, if flow into the port 98 is discontinued, by the natural inflow of water 14. A small through-flow of water leaks between the sleeve 50 and the seal assembly 103, and that overflow water, together with any bypass flow through the pressure relief valve 134 (Figure 2) passes to the high pressure side of the second seal 105. Again, there is a small water leakage past the seal 105, and that leakage flow together with any bypass flow through the pressure relief valve or valves passes in turn to the high pressure side of the seal assembly 107. At low pressures across the apparatus, only one of the seal assemblies 103, 105, 107 is energised, and which one it is is determined by natural leakage rates as described above.At pressures above the opening pressure of one of the relief valves, two seal assemblies out of the three are energised and the pressure drop.is shared between them. At the highest working pressures across the apparatus, both pressure relief valves are opened and all three seal assemblies 103, 105 and 107 share the pressure drop. An advantage of bringing seal assemblies 103, 105, 107 into use sequentially rather than simultaneously e.g.

by the use of pressure relief valves is that wear of the seal assemblies which are not energised is minimal. It will be of most benefit where the highest pressures requiring simultaneous energisation of all the seals occupy only a minor proportion of the duty time of the apparatus.

In the back-up seal housing 118, effective heat removal from the seal assembly 109 by thermal conduction is not so difficult to bring about because the seal housing 118 has a large surface area exposed to the air. The seal housing 118 is made from a strong metallic material such as aluminium bronze, and is bolted by retaining bolts 124 through the high intermediate and low pressure seal housings 112, 114 and 116 and directly into the shaft follower housing 66. Thus the back-up ring housing 118 has the mechanical strength to transmit the load from retaining nuts 119 at the ends of studs 124 to the high intermediate and low pressure seal housings 112, 114 and 116.In addition the high pressure seal ring 112 is independently bolted to the shaft follower housing 66, the intermediate pressure seal retaining ring 114 is bolted to the ring 112, and the ring 116 is independently bolted to the ring 114. The seal ring assembly 64 may therefore be dismantled in stages without total loss of the sealing function.

In use, effective sealing can be provided against ingress of fluid at the high pressure face of the assembly at up to about 13 bar per ring, the two pressure relief valves associated with high pressure seal assembly 103 and intermediate pressure seal assembly 105 causing the pressure to be shared substantially equally amongst the three energised series-connected seals 103, 105, 107. Concentricity of the seal assemblies 103, 105, 107, 109 and the shaft 44 is assured by the bearing 90. The leakage rate across the seal assembly 103, 105, 107 is substantially constant and is little affected by the pressure difference across the seal.

Preferably two pressure relief valves 134 as illustrated in Figure 2 are provided in each seal housing 112, 114. The valves are relatively small and typically have 4mm bores. Each is preset as described previously to lift at about 13 bar to assure that the pressure drop across each ring cannot exceed 10 to 12 bar even when the total pressure drop across the seal is in excess of e.g. 40 bar. The water leakage through the relief valve is expected to be very small, typically of the order of 2 - 5 litres per hour, since the pressure relief valves only have to handle the differences between the leakage rates past each sealing ring assembly, 103, 105 and 107, all of which are made to the same size and from the same material.

It will be appreciated that various modifications may be made to the embodiment described above without departing from the invention, the scope of which is defined in the appended claims. The pressure relief valves across the first and second seal ring assemblies could be replaced by pressure reducing valves or by a set of fixed orifices. The fluid used to energise seal assemblies 103, 105, 107 need not be the same, or of the same composition, as the fluid 14 being sealed against. Thus it would be possible to feed filtered or fresh water or other clean fluid to between the seal assemblies 103, 105 at pressure sufficient that there is a positive flow of filtered or fresh water to all of the seal assemblies 103, 105, 107, 109. Fresh or distilled water feed may be desirable where it is desired to minimise corrosion of the shaft 44 or sleeve 50.

Figure 3 shows part of the sealing ring 110 which is a narrow annular sleeve formed in two halves 110a and 110b. It is typically 20 mm wide and has a thickness typically 1% of the shaft diameter (e.g. 6 mm in the case of a shaft of diameter 600 mm). The ends of the half rings are lapped with a gap 140a, 140b between each pair of adjoining end faces. The gaps 140a, 140b are present to allow the ring halves 110a, 110b to expand or contract as wear takes place and they are usually about 1% of the shaft diameter. Thus for a 600 mm shaft the dimensions of the gaps 110a, 110b is about 5 to 6 mm.

A problem associated with the simple lap joint of Figure 3 is that the half sleeves 110a, 110b are free to move radially so that the gaps 140a, 140b at one junction close together leaving e.g. only 1 mm therebetween whilst the gaps 140a, 140b at the other junction have a gap of 10 mm. Development of such large gaps could allow fluid pressure behind the elastomeric ring 104 to force that ring into contact with the rotary shaft sleeve 50. This problem is minimised in the structure shown in Figure 4, where the lapped ends 142a, 142b of the half sleeves are formed with oppositely facing axially directed tongues 144a, 144b that interfit to define hooked joints that limit expansion of the gaps 140a and 140b and enable the half sleeves to be fitted initially with the gaps 140a, 140b at the diametrically opposed joints the same.

Instead of having two half sleeves 110a, 110b, the sealing ring 110 can be formed in one piece with a single split. However, a hooked joint as shown in Figure 4 is preferably still provided between the ends of the ring so as to minimise leakage, particularly under low pressure conditions.

A cross section of the sealing ring 110 appears in Figure 5 and it will be noted that the sides of the ring are slightly tapered. Provision of this taper stabilises the ring and minimises any tendency for the ring to rock against the face of the slot in which it sits.

It is preferred to make the ring 110 for large shafts from a low flexibility material such as a reinforced phenolic resin. Elastomeric materials such as reinforced nitrile or fluorocarbon rubber may also be used and conform more closely to the shaft at low applied pressures so that they have a lower initial leakage rate. However, for large diameter shafts a gap of about 2 mm is desirably provided between the seal housings 112 to 118 and the sleeve 50 and elastomeric materials are liable to extrude through that gap when high pressures appear across the seal.

The half sleeves 110a, 110b or a single split ring are therefore desirably manufactured in a relatively inflexible material and so that the joints are in close contact with the shaft surface when the sleeves or ring are initially fitted. If this is not done, the joints will tend to spring open and leakage can occur at the joints until there is sufficient pressure across the ring to close the joints down onto the shaft. For that purpose half sleeves 110a, 110b are manufactured with a curvature before pushing into the shaft 44 slightly greater than that of the sleeve 50 onto which they fit (see Figure 6).

Claims

1. Apparatus for sealing the space between the rotary shaft and the stationary member comprising at least two seals spaced axially apart along the shaft to be exposed to a higher pressure and a lower pressure, wherein means limits the pressure difference across the high pressure seal so that at pressures above the pressure limit both seals are active.

2. Apparatus according to Claim 1, wherein the pressure difference limiting means is connected between high pressure sides of the seals.

3. Apparatus according to Claim 1 or 2, wherein there are at least three seals connected in series via at least two pressure difference limiting means.

4. Apparatus according to Claim 1, 2 or 3, wherein the seals are housed in retaining rings of relatively high thermal conductivity.

5. Apparatus according to Claim 4, further comprising a seal on the low-pressure side that is not series-connected to the preceding rings.

6. Apparatus according to Claim 5, wherein the further seal on the low pressure side is connected at a high pressure side thereof to a drain passage so that abnormal flow through the drain passage indicates failure of preceding seals in the shaft.

7. Apparatus according to Claim 6, wherein the seal at the low pressure side is housed in a retaining ring of lower thermal conductivity but higher mechanical strength than the high thermal conductivity retaining rings.

8. Apparatus for sealing the space between the rotary shaft and a stationary member substantially as hereinbefore described with reference to and as illustrated in Figures 1 and 2, Figure 3 or 4, and Figure 5 of the accompanying drawings.

9. Apparatus for sealing the space between a rotary shaft and a stationary member surrounding the shaft, wherein the shaft is sealed by means of a plurality of seal elements and means controls the pressure exerted on each element so that at least at high pressure the pressure drop across the apparatus is distributed between the elements.

10. Apparatus according to Claim 9, wherein means is provided for regulating the pressure drop across each seal element.

11. Apparatus according to Claim 10, wherein the regulating means effects a simultanueous distribution of the pressure drop at all applied pressure drops.

12. Apparatus according to Claim 11, wherein the regulating means is at least one reducing valve or orifice connected across a seal element.

13. Apparatus according to Claim 9 or 10, wherein the regulating means energises the elements sequentially as the pressure across the apparatus increases.

14. Apparatus according to Claim 13, wherein the regulating means is at least one pressure relief valve connected across a seal element.

15. Apparatus for sealing the space between a rotary shaft and a stationary member surrounding the shaft, comprising a housing surrounding the shaft, a sealing ring contained in the housing, and means for applying fluid pressure to the sealing ring thereby to apply radially inward pressure to the sealing ring,,wherein the sealing ring has at least one split therein defined by a lapped joint and in its relaxed state has a curvature greater than that of the shaft.

16. Apparatus according to Claim 15, wherein the sealing ring is in a single part.

17. Apparatus according to Claim 15, wherein the sealing ring is in two or more radial segments.

18. Apparatus according to any of Claims 15 to 17, wherein the lapped portions of the or each joint are formed with inter-fitting hooks.

19. Apparatus according to any of Claims 15 to 18, wherein the sealing ring is formed of a material of low flexibility.

20. Apparatus according to Claim 19, wherein the sealing ring is formed of a reinforced phenolic plastics material.

21. Apparatus according to any of Claims 15 to 20, when the fluid pressure applied to the sealing ring is derived from the pressure of the fluid being sealed.

22. Apparatus according to any of Claims 15 to 21, wherein an elastomeric ring surrounds the sleeve, the fluid pressure being applied between the housing and the elastomeric ring and the elastomeric ring acting to transmit the radial pressure to the sleeve.

23. Apparatus according to Claim 22, wherein the sealing ring is contained in an annular space of the housing and is in sealing engagement with an axial face of the housing on the low pressure side of the sealing ring so that fluid can flow from a high pressure side of the housing to a space between the elastomeric ring and a wall of the housing spaced radially from the elastomeric ring, thereby to apply fluid pressure to the elastomeric ring.

24. Apparatus according to Claim 23, wherein the elastomeric ring fits between two axially spaced walls of the housing and is formed on its high pressure side with channels from which fluid can flow from the high pressure side of the sealing ring to the space between the elastomeric ring and the outer wall of the housing.

25. Apparatus according to any of Claims 15 to 24, wherein the fluid to be sealed against is water.