Abstract
A rotor for a fiberising apparatus includes a rotor housing; a first and second bearing assembly, wherein each bearing assembly has at least two ball bearings, each seated in a respective bearing seat; a substantially horizontal shaft rotatably mounted between the first bearing assembly and the second bearing assembly; and a plurality of resilient dampers arranged in an annular ring, wherein each resilient damper is releasably connected at a first end to the bearing seat and is releasably connected at a second end to the inner wall of the rotor housing.
Claims
1.-15. (canceled)
16. A rotor for a fiberising apparatus comprising: a rotor housing; a first and second bearing assembly, wherein each bearing assembly comprises at least two ball bearings, each seated in a respective bearing seat; a substantially horizontal shaft rotatably mounted between the first bearing assembly and the second bearing assembly; and a plurality of resilient dampers arranged in an annular ring, wherein each resilient damper is connected at a first end to the bearing seat and connected at a second end to the inner wall of the rotor housing.
17. The rotor according to claim 16, wherein each resilient damper is a frustum, or wherein each resilient damper is frustoconical.
18. The rotor according to claim 16, comprising a plurality of frustoconical resilient dampers.
19. The rotor according to claim 18, wherein each damper has a greater diameter at the damper face adjacent to the inner wall of the rotor housing and a lesser diameter at the damper face adjacent to the bearing seat.
20. The rotor according to claim 16, wherein the or each damper is any of a rubber damper, a silicone damper, a neoprene rubber damper.
21. The rotor according to claim 16, wherein the plurality of resilient dampers is adapted for a working rotation speed of the rotor of between about 4000 RPM and 13000 RPM.
22. The rotor according to claim 16, wherein the rotor housing has a greater wall thickness at the base of the rotor housing than at the upper surface of the rotor housing.
23. The rotor according to claim 16, wherein the rotor housing has a base wall thickness about 5 mm greater than the wall thickness at the upper surface of the rotor housing.
24. The rotor according to claim 16, wherein the internal profile of the rotor housing is asymmetric.
25. The rotor according to claim 16, wherein the bearing seat is substantially cylindrical, the rotor housing is substantially cylindrical and the central axis of the bearing seat is offset from the central axis of the rotor housing.
26. The rotor according to claim 16, wherein the or each damper has a Shore A hardness of about 55.
27. The rotor according to claim 16 wherein the damper ring stiffness is between about 5.Math.10.sup.5 to about 10.sup.6 N/m.
28. The rotor according to claim 16, wherein each damper comprises a threaded screw for releasable connection to the bearing seat.
29. The rotor according to claim 16, wherein each damper comprises a threaded aperture for releasable connection with a screw through the rotor housing.
30. The rotor according to claim 16, comprising between about 10 and about 24 frustoconical dampers arranged annularly.
31. The rotor according to claim 16, wherein the dampers are equidistant from each other around a substantially annular bearing assembly.
32. The rotor according to claim 16, wherein the or each bearing is a hybrid angular ball bearing.
33. The rotor according to claim 16, wherein the inner diameter of the or each ball bearing is between about 60 mm and about 75 mm.
34. The rotor according to claim 16, wherein the bearing is made of a ceramic material.
35. The rotor according to claim 16, wherein the or each bearing is a hybrid angular ball bearing the distance between the two angular contact bearings is about 20 mm.
36. The rotor according to claim 16, wherein the clearance between the annular bearing seat and the inner surface of the rotor housing is about 14 mm.
37. The rotor according to claim 16, wherein the or each bearing is a hybrid angular ball bearing and wherein the two angular ball bearings are spaced apart by an inner axial spacer ring and an outer axial spacer ring; wherein the width of the outer spacer ring is less than the width of the inner spacer ring.
38. The rotor according to claim 16, wherein the or each bearing is a hybrid angular ball bearing wherein the two angular ball bearings are spaced apart by an inner axial spacer ring and an outer axial spacer ring and wherein the width of the outer spacer ring is between about 16 μm and 61 μm less than the width of the inner spacer ring.
39. The rotor according to claim 16, wherein the relationship between the shaft diameter (D.sub.shaft) and the shaft length (L.sub.shaft) is defined as: D.sub.shaft(L.sub.shaft)≥0.12*L.sub.shaft−32 mm for a range of shaft lengths between about 101 mm and 1325 mm and for a range of shaft diameters greater than or equal to 20 mm and for a seat stiffness (damper ring stiffness) of less than or equal to 3*10.sup.6 N/m.
40. The rotor according to claim 16, wherein the outer cross-sectional diameter of the shaft is about 100 mm; the shaft has a bearing seat diameter of about 70 mm; the length of the shaft is about 955 mm; and wherein the length of the shaft between a centre-point of the first bearing assembly and a centre-point of the second bearing assembly is about 590 mm.
41. The rotor according to claim 16, wherein the weight of the or each bearing seat is less than or equal to 3 kg.
42. A fiberising apparatus comprising a set of at least three rotors according to claim 16, wherein each rotor is mounted for rotation about a different substantially horizontal axis and arranged such that, when the rotors are rotating, melt poured on to the periphery of the first rotor in the set is thrown successively onto the periphery of each of the subsequent rotors and fibres are thrown off from the rotors.
43. A method of manufacture of man-made vitreous fibres (MMVF) comprising: providing a fiberising apparatus comprising a set of at least three rotors according to claim 16, each mounted for rotation about a different substantially horizontal axis, wherein each rotor has a drive means; rotating the rotors; providing a mineral melt for formation of man-made vitreous fibres (MMVF) wherein the melt is poured on to the periphery of the first rotor; collecting the fibres formed.
Description
[0098] The invention will now be described by way of example with reference to the accompanying drawings, in which:—
[0099] FIG. 1 is a longitudinal cross-sectional view of a rotor according to the present invention;
[0100] FIG. 2 is a perspective view of a right position spinner comprising four rotors in accordance with the present invention, shown without the spinner housing;
[0101] FIG. 3 is a perspective view of a bearing assembly at the non-drive end (NDE) of the rotor of FIG. 1;
[0102] FIG. 4a is a perspective view of the bearing seat of the bearing assembly of the rotor in accordance with the present invention without the dampers shown;
[0103] FIG. 4b is a perspective view of a section of the bearing seat shown in FIG. 4a;
[0104] FIG. 5a is a cross-sectional view through a frustoconical damper shown in the bearing assembly of FIG. 3;
[0105] FIG. 5b is a perspective view of a cross-section of the frustoconical damper of FIG. 5a;
[0106] FIG. 5c is a perspective view of the damper of FIGS. 5a and 5b;
[0107] FIG. 6a is a perspective view of an alternative embodiment of a bearing assembly at the non-drive end (NDE) of a rotor in accordance with the present invention;
[0108] FIG. 6b is a perspective view of a section of the bearing seat shown in FIG. 6a;
[0109] FIG. 7 is a cross-sectional view through a bearing assembly at the non-drive end (NDE) of the rotor showing the greater clearance between the bearing assembly and the rotor housing;
[0110] FIG. 8a is a perspective view of the rotor of the present invention showing half of the rotor housing removed;
[0111] FIG. 8b is a plan view of the NDE of the rotor housing, bearing seat and dampers of the present invention;
[0112] FIGS. 9a, 9b and 9c are schematic cross-sectional views (not to scale) through a pair of angular contact ball bearings in the rotor of the present invention, exemplifying the preferred reduction in bearing preloads due to temperature differences, wherein FIG. 9b shows the shaft when cold and FIG. 9c shows the shaft when hot;
[0113] FIG. 10 is an external perspective view of the rotor of the present invention showing the cooling and air purging systems;
[0114] FIG. 11 is a cross-sectional view of the DE bearing seat showing a grease labyrinth;
[0115] FIG. 12 is a front view of the right position spinner of the present invention showing all four rotors with mounted wheels; and
[0116] FIG. 13 is a rear view of a 5D surface and contour plot of rotor 4 simulating the lifetime for the NDE bearing seat for varying shaft diameter, mass of seat, shaft length and seat rubber stiffness (damper ring stiffness) at a maximum rotor speed of 13000 RPM;
[0117] FIG. 14 shows a view from above of the 5D surface and contour plot of FIG. 13, showing mass of seat (M.sub.seat kg), shaft diameter of middle section (D.sub.shaft mm), shaft length between seat centres (L.sub.shaft mm) and seat rubber stiffness (K.sub.seat N/m); and
[0118] FIG. 15 shows a 5D and contour plot of rotor 4 simulating the displacement of the NDE seat (Δx.sub.seatNDE) at a maximum rotor speed of 13000 RPM.
[0119] Referring to FIG. 1, a longitudinal cross-section of the rotor 1 is shown, with the shaft 2 longitudinally positioned between a drive end (DE) 3 and non-drive end (NDE) 4 of the rotor 1 to be substantially horizontal. The shaft 2 is a hollow cylindrical steel shaft with an outer diameter of about 100 mm and a bearing seat diameter of about 70 mm. In alternative embodiments, the shaft outer diameter is between about 100 mm and about 120 mm and the bearing seat diameter is between about 50 mm and about 100 mm. The length of the shaft 2 between a centre-point of the bearing at the DE 3 and a centre-point of the bearing at the NDE 4 is between about 530 mm and 590 mm. For the embodiment shown, the length of the shaft 2 is about 590 mm. At the NDE 4 the shaft 2 “over-hangs” beyond the first bearing assembly 5 and the total length of the shaft 2 is about 955 mm. A second bearing assembly 6 is positioned at the DE 3. The bearing assembly 5, 6 for both the NDE 4 and the DE 3 are identical (except for minor details like cooling water supply to the wheel), but the NDE 4 bearing assembly 5 is described in more detail with respect to FIGS. 3 to 5. The first and second bearing assembly 5, 6 are each placed adjacent to a “soft” suspension, with the respective first and second soft suspensions acting independently of each other. The NDE 4 may comprise more than one bearing assembly 5, which will increase bearing lifetime as the static load on each bearing assembly 5 is reduced.
[0120] Referring to FIG. 2, a perspective view of the four rotors of a right position spinner is shown without the spinner body; a first rotor 1a, a second rotor 1b, a third rotor is and a fourth rotor 1d, are shown. Right position refers to the position of fourth rotor 1d to the right. A variant (not shown) is a left position spinner, which is a mirror image of the right position spinner shown in FIG. 2, but with the fourth rotor 1d positioned to the left. The first rotor 1a has a maximum speed of 6000 RPM and is connected to a first motor 30a. The second rotor 1b has a maximum speed range of between 6000-13000 RPM and is connected to a second motor 30b. The third rotor is has a maximum speed range of between 6000-13000 RPM and is connected to a third motor 30c. The fourth rotor 1d has a maximum speed range of between 6000-13000 RPM and is connected to a fourth motor 30d. The wheels 13b, 13c, 13d attached respectively to the second, third and fourth rotors are shown. Each motor 30a, 30b, 30c, 30d and the respective motor shaft are fixed at the spinner body. The rotor shaft 2, shown in FIG. 1, is mounted flexibly inside the rotor housing 12, which is also fixed. A flexible lamella coupling 32 connects the shafts of each motor drive end and drive end of each rotor and allows for some radial misalignment. However, the maximal allowable radial misalignment of the present coupling is 1.3 mm. Hence the coupling puts a restriction on the allowable displacement of the shaft 2 at the DE 3. With reference to FIG. 8a, there are further connections between the motors 30a, 30b, 30c, 30d and the rotors 1, 1b, 1c, 1d to power and cool the spinner.
[0121] With reference to FIGS. 3, 4a and 4b the bearing assembly 5 is shown in more detail and comprises two ball bearings 7 held within an annular bearing seat 8. The bearing seat 8 has a stainless-steel body shown in FIG. 4a and FIG. 4b having a plurality of recesses 9a equally spaced around its outer surface with damper seats 9b equally spaced therebetween. Referring to FIG. 4b, each damper seat 9b has a threaded hole 9c into which a damper 11 is secured. As shown in FIG. 4b, the recesses 9a are truncated cylindrical recesses. The bearing seat 8 has a very low mass of 3 kg with the minimum of material to support the components held by the bearing seat 8. The bearing seat 8 also comprises an aluminium labyrinth ring 10 to reduce the seat mass.
[0122] With reference to FIG. 3 and FIG. 4a, each damper seat 9b supports one end of a frustoconical damper 11, which protrudes from the bearing seat 9b and the bearing seat 8. As shown in FIGS. 5a, 5b and 5c, the dampers 11 are frustoconical with a threaded metal screw 11a at a first smaller inner end face or tap end. The inner end face has a diameter of about 20 mm. There is a further metal insert 11b at the second outer end face or hole end, which has a central threaded aperture 11c partially into the rubber damper 11. The outer end face has a diameter of about 27 mm and the length of the damper 11 is about 25 mm. With reference to FIGS. 3 and 4b, the damper 11 is connected to the bearing seat 8 by a screw connection, whereby the threaded screw 11a of the damper 11 engages a threaded hole 9c in the bearing seat 8. Similarly, the outer end face of the damper 11 is provided with the threaded aperture 11c for engagement with a threaded screw or bolt for connecting the damper 11 to the rotor housing (not shown). The dampers 11 are made of neoprene rubber with a Shore A hardness of about 55 and a damper ring stiffness of about 10.sup.6 N/m. It is understood that, although the embodiment shown is for a bearing assembly/suspension at the NDE 4, the damper ring stiffness of about 10.sup.6 N/m is substantially similar at the DE 3. In alternative embodiments, the dampers are cylindrical, but in the preferred embodiment shown in FIG. 3, for rotor 4, there are twenty frustoconical dampers 11 spaced equidistant from each other around the bearing seat 8. Each damper 11 does not touch the adjacent damper 11 in the annular arrangement. The embodiment shown in FIG. 3 is for the fourth rotor of a fiberising apparatus having four rotors. In alternative embodiments, as shown in FIG. 6a, the rotor has between about twelve and about twenty-two frustoconical dampers 11′ equidistant from each other.
[0123] FIG. 6a and FIG. 6b show an alternative embodiment of the first bearing assembly 5′ and a bearing seat 8′ for the third and fourth rotors having twelve substantially cylindrical dampers 11 spaced equidistant from each other around the bearing seat 8′, with each damper 11 having a Shore A hardness of about 55 and damper ring stiffness of about 10.sup.6 N/m. The damper 11 is connected to the bearing seat 8′ by a screw connection, in that a threaded screw of the damper 11 is engaging a threaded hole in the bearing seat 8′. Similarly, the other end of the damper 11 is provided with a threaded aperture for engagement with a bolt for connecting the damper 11 to the rotor housing. The bearing seat 8′ further comprises a ball bearing inner steel ring 16′ and aluminium labyrinth ring 10′, which is used to reduce the seat mass.
[0124] Referring to FIGS. 1, 2 and 7, the dampers 11 are positioned between the bearing seat 8 and the rotor housing 12 to form an annular bearing ring. As shown in FIG. 7, the rotor housing 12 of the present invention has a clearance of about 14 mm between the bearing assembly 5 and the inner surface of the housing 12, both at the NDE 4 and the DE (not shown) In use, the rotor 1 would be attached to a high strength steel wheel at a position 13 further along the shaft 2 from the first bearing assembly 5 at the NDE 4. Referring to FIG. 2, the wheel 13a, 13b, 13c, 13d is placed at a first end of the shaft 2 and the opposite end is connected to a high-speed motor 30a, 30b, 30c, 30d by a flexible coupling. There is a relatively narrow space between the wheels 13a, 13b, 13c, 13d. For the fourth rotor example shown in FIG. 1, the wheel is substantially cylindrical with a weight of about 50 kg, a diameter of about 332 mm, an outer wall thickness of about 25 mm and a side wall thickness of around 15 mm. It is understood that the outer wall of the wheel is a curved surface and the side walls are circular and substantially perpendicular to the length of the shaft 2. With reference to FIG. 2; typically, the wheel diameter of wheel 3 is about 314 mm with a mass of about 50 kg. For a preferred embodiment of the fourth rotor, the wheel is about 47 mm from the centre line of the NDE bearing assembly 5 and rotates at between about 6000 RPM and about 13000 RPM. The rotor housing 12 supports the bearings and covers the middle part of the shaft 2. However, for ease of understanding of the invention the wheel is not shown in FIG. 1.
[0125] Referring to FIGS. 2 and 8a, each rotor housing 12 comprises two symmetrical mating parts, each of which is a semi-cylindrical shell, such that when the two parts mate a substantially cylindrical housing 12 is formed. Each half of the rotor housing 12 is held to the adjacent half of the housing by screws 33. FIG. 8a also shows the screws 11d holding each damper (not shown) to the rotor housing. The rotor housing 12 further comprises brackets 35 that support pipes and wires. A rubber ring 34 at each end 4, 5 of the rotor housing 12 is used for mounting the rotor 1 in the spinner.
[0126] Referring to FIG. 8b, in preferred embodiments, the rotor housing 12 has a greater wall thickness of about 5 mm at the base of the housing 12 compared to the upper wall thickness of the rotor housing 12, which effectively lifts the wheel to compensate for the overhanging effect. The internal profile of the rotor housing 12 is asymmetric. Referring to FIG. 8b, the dampers 11′ in the top part of the NDE 4 are stretched by about 2.5 mm. The dampers 11″ in the bottom part of the NDE 4 are compressed by about 2.5 mm. The offset arrangement of the central axis of the annular mounting ring/bearing seat 8 with respect to the central axis of the annular rotor housing 12 counteracts the effect of gravity on the suspension of the bearings 7, shaft 2 and wheel 13 in the rotor housing 12 so that the centre of gravity of the rotor 1 is in the desired position.
[0127] With reference to FIGS. 3 and 6b, the inner surface of each annular bearing seat 8 supports two hybrid ball bearings 7, which are lifetime lubricated super precision hybrid angular contact ball bearings 7 that have been selected to achieve the desired speed and lifetime before failure. The two bearings 7 at each of the DE 3 and NDE 4 are fitted close together with little space between them at each end of the rotor 1.
[0128] Referring to FIGS. 9a, 9b and 9c, the bearings are ceramic ball bearings 7 between two steel bearing rings 16, 17. The bearings 7 have been selected to withstand loads of up to 1400N. The dynamic load unbalance at the wheel of rotor 4 was determined to be about 560 g.Math.cm and the configuration of the bearing seats and the damper suspensions was optimised to compensate for this unbalance. The angular contact ball bearings 7 have a diameter of about 70 mm, but in alternative embodiments the bearing diameter is any of 60 mm, 65 mm, 70 mm or 75 mm.
[0129] In the embodiment of FIGS. 9a, 9b and 9c the ball bearing 7 has a diameter of 70 mm. Each of the pair of ball bearings 7 are axially spaced apart by an inner spacer ring 14 between the rotating inner surfaces of the ball bearings 7 adjacent to the shaft 2 and an outer spacer ring 15 between the rotating inner surfaces of the ball bearings 7 furthest from the shaft 2. The ball bearings 7 are held between a ball bearing inner steel ring 16 mounted on the steel shaft 2 and a ball bearing outer steel ring 17. The angular contact ball bearings 7 each have an angular contact (pressure) angle β, which is symmetrical about the centre line between the ball bearings 7. The contact angle β is the pressure direction during rotation. The contact angle β will vary depending on the temperature of the steel shaft 2.
[0130] Referring to FIG. 9b, the contact angle β will increase if the shaft 2 is colder because the shaft will contract to a smaller diameter. Referring to FIG. 9c, the contact angle β decreases if the shaft 2 is hotter because the shaft will expand to a greater diameter. As shown in FIG. 9b when the shaft 2 is cold it has a lesser diameter R.sub.c and the pressure angle β is greater. The outer spacer ring 15 has a width W.sub.c and does not make contact with both the ball bearing inner steel ring 16 and the ball bearing outer steel ring 17 and there is greater scope for movement/rattling of the ball bearings 7. However, as shown in FIG. 9c, when the shaft 2 is hot it has a greater diameter R.sub.H and the pressure angle β is less. The increased diameter of the shaft also means that the balls 7 of the bearing push the outer bearing seats 17 towards each other until then are in contact with the outer spacer ring 15. Thus, the bearing assembly 5, 6 has been arranged to reduce the bearing preloads due to the expected temperature difference when the rotor is in use, which is about 10° C.
[0131] Referring to FIG. 9b and FIG. 9c, it has been found that it is advantageous that the outer spacer ring 15 is shorter than the inner spacer ring 14. Typically, for a bearing diameter of 70 mm for rotor 3 and 4, the width of the outer spacer ring 15 is about 61 μm less than the inner spacer ring 14. For rotor 2 for a bearing diameter of 70 mm, the width of the outer spacer ring 15 is about 16 μm less than the inner spacer ring 14.
[0132] As shown in FIGS. 3 and 10, the annular bearing seat 8 is surrounded by end rings 18.
[0133] As shown in FIG. 10, the rotor 1 has a cooling system for both the DE bearing assembly 6 and the NDE bearing assembly 5. An NDE cooling fluid inlet 20 and an NDE cooling fluid outlet 21 are connected by channels through which cooling fluid flows to carry heat away from the NDE 4 of the rotor 1. A DE cooling fluid inlet 23 and DE cooling fluid outlet 24 are also connected by channels to carry heat away from the DE 3 of the rotor 1. A fluid inlet 25 for water cooling of the wheel (not shown) is also provided. In the embodiment shown the cooling fluid is water. It has been shown through testing that the bearing temperatures can all be maintained at about 50° C.
[0134] Referring to FIG. 10, the rotor 1 further comprises a DE air purge inlet 26 and an NDE air purge inlet 27. The air purge system flushes contaminants from the rotor 1. The rotor 1 further comprises an accelerometer 28 at the NDE 4 and an accelerometer 29 at the DE 3 and a pyrometer 30 at the DE 3 and a pyrometer 31 at the NDE 4.
[0135] Referring to FIGS. 10 and 11, the air purging system flushes contaminants from the rotors when the spinner is in use. The rotor also comprises a water-cooling system with a cooling water outlet 24 to carry heated water away from the bearings 7. It has been found that, when washing the spinner during regular servicing, there is potential for water to contaminate the bearings 6. Thus, the present invention further comprises a grease labyrinth 38, as shown in FIG. 11, to prevent water polluting the bearings 7 during washing of the spinner. FIG. 11 shows the DE 4 bearing assembly 6 comprising a rubber ring 34 for mounting the rotor 1 in a spinner. For the DE 3 shown, there is a DE labyrinth grease inlet 38a and an NDE labyrinth grease inlet 38b. The NDE labyrinth grease inlet 38b carries grease to the NDE 4, which has a similar grease labyrinth. When the rotor is still because the spinner is being serviced, the labyrinth 38 is sealed with grease to prevent water polluting the bearings 7 during washing. Sealing rings 39 prevent labyrinth grease from flowing into the bearings 7, which are lifetime lubricated with their own grease.
[0136] Referring to FIG. 12, in use, a fiberising apparatus comprising four rotors 1a, 1b, 1c, 1d is used to manufacture man-made vitreous fibres (MMVF).
[0137] Each rotor 1a, 1b, 1c, 1d is mounted for rotation about a different substantially horizontal axis and each rotor 1a, 1b, 1c, 1d has a drive means, which may be a single drive means to power all four rotors 1. In a preferred embodiment, a second rotor comprises a suspension ring at the non-drive end bearing assembly having fourteen neoprene rubber dampers and the third and fourth rotor at the non-drive end have a bearing assembly with twenty neoprene rubber dampers. The second rotor comprises twelve neoprene rubber dampers at the drive end bearing assembly and the third and fourth rotors have eighteen neoprene rubber dampers at the drive end bearing assembly.
[0138] Each right position rotor 1a, 1b, 1c, 1d has a high-speed motor connected at one end by a flexible coupling and a wheel is placed at the opposing end of the rotor 1a, 1b, 1c, 1d. In the embodiment shown in FIG. 12, rotors 1 and 3 rotate in an anticlockwise direction and rotors 2 and 4 rotate in a clockwise direction. The wheel of the first rotor 1a has a diameter of about 184 mm; the wheel of the second rotor 1b has a diameter of about 234 mm; the wheel of the third rotor 1c has a diameter of about 314 mm and the wheel of the fourth rotor 1d has a diameter of about 332 mm. The space between the outer surface of each mounted wheel varies, with the greatest distance of about 228 mm between the wheels of the first and fourth rotor 1a, 1d and the least distance of about 17 mm between the first and second rotors 1a, 1b.
[0139] Referring to FIG. 12, with the rotors 1a, 1b, 1c, 1d rotating a molten mineral melt of stone or rock or a slag or glass melt is poured through an inlet 36 onto the periphery of the wheel of the first rotor 1a to spin the melt and throw off MMVF. The melt is then thrown successively to the wheels of the remaining three rotors 1b, 1c, 1d and fibres are formed and collected each time. Simultaneously, as the melt passes to the wheel of each successive rotor 1a, 1b, 1c, 1d, a high-pressure air flow through the spinner and along the wheel removes fibres from the wheels for collection.
[0140] For a typical four rotor fiberising apparatus according to the present invention, wheel 1 produces about 5% of the hourly stone wool production; wheel 2 produces about 25%, wheel 3 produces about 40% and wheel 4 produces about 30%. Manufacture using the rotors of the present invention can continue for around 4000 hours before the ball bearings 7 require replacement, which is a significant increase over known devices. An endurance test comparing a prior art spinner running at max speed (9300 RPM) with a spinner according to the present invention running at 13000 RPM, both operating with an unbalance of 560 g.Math.cm, showed an improvement from 603 hours to more than 4000 hours. Further tests have found a mean time between failure of approximately 15000 hours for a spinner according to the present invention.
[0141] Referring to FIG. 13, to evaluate the optimum parameters for the improved rotor of the present invention, simulations were conducted for a range of shaft lengths (L.sub.shaft), shaft diameter (D.sub.shaft), damper ring stiffness (K.sub.seat), bearing seat mass (M.sub.seat) and bearing diameter. The effect of these values on the bearing lifetime (L.sub.10) and rotor shaft displacement (Δx) were shown in 5D plots.
[0142] FIG. 13 is a rear view of a 5D contour plot for rotor 4 at a maximum rotor speed of 13000 RPM for an NDE seat having a bearing diameter of 70 mm. The height of each peak (L.sub.10) represents the bearing lifetime in hours or the Basic Life Rating. The bearing lifetime statistic used is a measure of the amount of time in revolutions when 90% of the ball bearings can be expected to survive.
[0143] As shown in FIG. 13, the optimum configuration of the present invention allows for a maximised bearing lifetime of 3617 hours. The simulation considers a range of shaft diameter (D.sub.shaft) and bearing seat mass (M.sub.seat) plotted together with the shaft length (L.sub.shaft) between bearing seat centres and rubber stiffness (K.sub.seat) at the bearing seat. Each cube shown in the 5D plot of FIG. 13 symbolises a 3D plot of the function of the lifetime (L.sub.10) with shaft length (L.sub.shaft) and shaft diameters (D.sub.shaft) as arguments. In each 3D subplot are rubber stiffness (K.sub.seat) and bearing seat mass (M.sub.seat) constants as dictated by the position of the 3D subplot.
[0144] FIG. 14 shows a view from above of the 5D reface and contour plot of rotor 4 shown in FIG. 13. Each square of the plot shows a shaft length range of 101 to 1325 mm in increments of 101 mm, 407 mm, 713 mm, 1019 mm and 1325 mm and a shaft diameter range of 20 to 170 mm in increments of 20 mm, 58 mm, 95 mm, 133 mm and 170 mm. The range of rubber stiffness (K.sub.seat) is from 10.sup.4N/m to 10.sup.8 N/m. and bearing seat mass (M.sub.seat) from 1.5 to 12.0 kg.
[0145] Referring to FIG. 13 and FIG. 14, the rubber stiffness (K.sub.seat) at the bearing seat is optimised at about 10.sup.6 N/m or lower. It is also preferable that the bearing seat mass (M.sub.seat) is as low as possible, with improvement in the bearing lifetime for the reduced 3 kg mass of the bearing seat of the present invention. It was also found that if the seat mass (M.sub.seat) is reduced too far; for example, to about 1.5 kg it would be necessary to reduce the rubber stiffness (K.sub.seat) at the bearing seat to achieve any gain in bearing lifetime. It is also shown that for the shaft length (L.sub.shaft) of 590 mm and shaft diameter (D.sub.shaft) of 100 mm for the present invention, an improvement in bearing lifetime is achieved. Further embodiments of the present invention, with a bearing diameter of 60 mm have been investigated and it was found that the present invention can achieve a maximum bearing lifetime of 6183 hours.
[0146] Referring to FIG. 15, a 5D surface and contour plot for rotor 4 is shown plotting displacement of the NDE seat (Δx.sub.seatNDE) at maximum rotor speed of 13000 RPM for an NDE seat having a bearing diameter of 70 mm. As shown, for the maximum allowable displacement (Δx.sub.seatNDEmax) of 15 mm the seat rubber stiffness is 10.sup.5 N/m. Thus, it was concluded that, to avoid exceeding the maximum allowable bearing seat displacement, the seat rubber stiffness (damper ring stiffness) should be greater than 10.sup.5 N/m.
[0147] Within this specification, the term “about” means plus or minus 20%; more preferably, plus or minus 10%; even more preferably, plus or minus 5%; most preferably, plus or minus 2%.
[0148] Within this specification, the term “substantially” means a deviation of plus or minus 20%; more preferably, plus or minus 10%; even more preferably, plus or minus 5%; most preferably, plus or minus 2%.
[0149] The above described embodiment has been given by way of example only, and the skilled reader will naturally appreciate that many variations could be made thereto without departing from the scope of the claims.
