rotor

Abstract

To relate to a rotor for a yarn collecting device for use in the manufacturing of an artificial vitreous fiber (MMVF), and a manufacturing method of the artificial vitreous fiber (MMVF).SOLUTION: A rotor for a yarn collecting device comprises: a rotor housing; first and second bearing assemblies each having at least two ball bearings which are seated in each bearing seat; and a substantially-horizontal shaft rotatably arranged between the first bearing assembly and the second bearing assembly. A plurality of elastic dampers are arranged at an annular ring, and each elastic damper is connected to the bearing seat at a first end part so as to be releasable, and connected to an inner wall of the rotor housing at a second end part so as to be releasable.SELECTED DRAWING: Figure 1

Description

[Technical field]

[0001] The present invention relates to an improved rotor, in particular to a rotor for a yarn harvester for use in the manufacture of man-made vitreous fibers (MMVF) and to a method of manufacture of man-made vitreous fibers (MMVF). [Background technology]

[0002] Spinning devices known as spinners or (cascade) spinners are used in the manufacture of MMVFs for producing insulating materials, e.g., to provide soundproofing or thermal insulation materials from mineral melts of stone or rock or slag or glass melts. The spinners have a set of rotors for spinning molten material or lava onto a spinning wheel to produce a web of insulating product. Molten stone or lava ("melt") is fed successively from a first rotor to the remaining rotors of the set, with fibers being discharged from each wheel as each of the rotors rotates. These fibers are collected and removed from the set of rotors for the production of insulating products, such as stone wool insulating products.

[0003] The rotors in the spinner operate at very high speeds. By controlling the force of the high speed and high acceleration of the spinner, the physical and performance properties of the fiber and therefore the insulation produced are controlled. It has been found that by increasing the speed and acceleration of the spinning device, the spun fiber can be made finer and softer with improved and highly desirable insulating properties. It has been found that with finer fibers, there is less conduction in the spun fiber and more air is retained in the insulation product when the insulation product is made from finer fibers.

[0004] Known spinners operate at high speeds and accelerations of about 150 km / s2 to achieve very fine fibers required for good thermal insulation properties. Each rotor wheel comprises a rotating shaft suspended between bearings at each of the drive end (DE) and non-drive end (NDE). Because the shaft at the NDE passes over the rotor to the wheel where the molten material is directed, the NDE and DE of the rotor are not equidistant from each end of the shaft. The NDE of the shaft effectively overhangs the bearings and the NDE of the rotor has been found to have the highest load. Vibrations of the rotor mechanics at the DE and NDE during spinning cause significant wear to the bearings seated at each end of the rotor and wear to dampers located between the rotor housing and the spinner body. Known devices use dampers located between the rotor housing and the spinner body to reduce vibrations passed from one rotor to the other in the set.

[0005] It is known to use springs as vibration absorbing means. US Patent No. 2,556,317 discloses a bearing assembly for a centrifuge in which a radial compression spring or rubber cushion is arranged radially between the bearing element and the stationary frame of the machine. WO 2014 / 000799 discloses a spring damping element for an electric compressor / turbine generator. The damping element is a spring steel ring having a recess in which a leaf spring is received to apply a radial force.

[0006] However, the scale and speed of spinning in the stone wool harvesting equipment places very high loads on the bearings used, resulting in the need to frequently replace worn bearings.Typically, a four-wheel spinner in which the rotor of the present invention is used produces 5-6 tons of stone wool per hour, and the resulting reduction in "downtime" due to maintenance significantly increases the volume of product that can be produced.

[0007] The rotors in the yarn harvesting spinner are each arranged about a substantially horizontal axis, which results in wear on the rotor bearings as gravity promotes imbalance, which leads to variable wear on the bearings. It has also been found that imbalances in the forces applied to each rotor are caused by the melt being dumped onto the rotor wheels. The imbalance is exacerbated if there is uneven wear of the rotor mechanics or wear of the outer surface of the rotor assembly through which the molten material is conducted. One example is the deposition of a layer of solidified melt on the rotor, so-called "freeze lining", which can be uneven and can detach in areas that cause the imbalance. The combination of these factors causes the bearings used in existing spinners to wear out much faster than desired, requiring the spinner to be taken out of production for maintenance. There is therefore a significant need to improve rotor configurations and dynamics to increase the mean time between failures. Summary of the Invention [Problem to be solved by the invention]

[0008] The present invention aims to provide an improved rotor for a thread harvesting device which addresses the above mentioned problems associated with imbalance and high rotational forces which cause wear on the rotor bearings. [Means for solving the problem]

[0009] In a first aspect, the present invention provides a rotor for a thread harvesting device comprising a rotor housing, first and second bearing assemblies, each bearing assembly comprising at least two ball bearings each seated in a respective bearing seat, a substantially horizontal shaft rotatably mounted between the first and second bearing assemblies, and a plurality of elastic dampers arranged in an annular ring, each elastic damper being releasably connected at a first end to the bearing seat and at a second end to an inner wall of the rotor housing.

[0010] It will be understood that in the context of the present invention, "rotor" is understood to refer to the rotating assembly, "shaft" is understood to refer to the long cylindrical rotating rod used to transmit rotational power, "damper" is understood to be a device for suppressing or absorbing vibrations, and an "inner wall" of the rotor housing is understood to be the wall facing towards the shaft of the rotor.

[0011] Preferably, each resilient damper is a frustum, more preferably, each resilient damper is a frustum of a cone.

[0012] The present invention reduces wear on the bearings, so that the rotor can withstand the high speeds and significant loads applied by the spinner. The annular arrangement of multiple elastic dampers is particularly well suited for the instabilities that occur in the high speed spinner of the present invention. The present invention is a significant improvement over known solutions used in turbines. For example, elastic ring-shaped springs or dampers do not protect the bearings sufficiently to achieve the improved mean time between failures achieved by the present invention. The "soft" suspension of the present invention significantly increases the mean time between failures of the bearings / rotor, thereby improving the efficiency of the silk harvesting device by reducing maintenance time. By releasably connecting the elastic dampers to the bearing seats and to the inner wall of the rotor housing, the dampers work in both compression and tension, resulting in a significant increase in the mean time between failures of the bearings. The present invention avoids potential problems with the natural frequency of the damper and avoids loss of action that can occur when using springs, for example if the spring loses contact at one end.

[0013] It has been found that the use of multiple frusto-conical dampers can significantly reduce the load on the bearings, even when the rotor speed reaches a critical speed for the bearings. As an example, testing at 13,500 RPM with and without the soft suspension of the present invention showed a reduction in dynamic load from 2760N to 192N.

[0014] The multiple frustoconical dampers minimize the internal wear of the bearing by providing improved absorption of vibrations generated by high speed / high acceleration and imbalance of the spinning rotor. In effect, each damper is arranged and shaped to provide more elastic material where it is most needed to withstand static and dynamic loads applied to the bearing assembly. It has been demonstrated that frustum, cone or frustoconical dampers better withstand static and dynamic loads in the bearing suspension. The imbalances that occur when the shaft is rotating at high speed, for example when the melt is poured or the fibers are discharged, are absorbed by the elastic dampers that expand or contract according to the forces applied to the bearing assembly. The solution of the present invention is particularly suitable for use with a yarn harvesting device and is carefully configured to optimize the volume and stiffness of the rubber placed in the annular ring and extend the life of the bearing. It has been found that the optimal choice is to maximize the rubber volume within the constraints of the spinner while making the suspension as "soft" as possible.

[0015] Optionally, each elastic damper is cylindrical. Alternatively, each elastic damper is a frustum of a pyramid. However, preferably, each elastic damper is rotationally symmetric.

[0016] The ease of mounting of the dampers in the bearing seats of the spinner is improved by the dampers having rotational symmetry, i.e. they are rotationally symmetric about their central axis.

[0017] Preferably, the rotor includes a plurality of frusto-conical resilient dampers forming an annular ring between the bearing seat and the inner wall of the rotor housing.

[0018] Preferably, the rotor comprises a plurality of frusto-conical resilient dampers, each having a larger diameter at the inner wall of the rotor housing and a smaller diameter at the bearing seat.

[0019] Preferably, each damper comprises a threaded screw for releasable connection to the bearing seat, and / or each damper comprises a threaded opening through the rotor housing for releasable connection with a screw.

[0020] Preferably, the rotor housing further comprises at least one threaded screw receivable by a threaded opening in the damper.

[0021] The releasable connection of each damper allows for quick and convenient replacement, improving the efficiency of rotor maintenance.

[0022] Preferably, the rotor housing has a greater wall thickness at the base of the rotor housing than at the top surface of the rotor housing.

[0023] Preferably, the rotor housing has an increased base wall thickness and a decreased top wall thickness, in that the base wall thickness is increased by about 2 mm to about 3 mm and the top wall thickness of the rotor housing is correspondingly decreased, more preferably the rotor housing has an increased base wall thickness by about 2.2 mm to about 2.7 mm and the top wall thickness of the rotor housing is correspondingly decreased, and most preferably the rotor housing has an increased base wall thickness of about 2.5 mm and a reduced top wall thickness of about 2.5 mm when compared to the standard wall thickness of the rotor housing.

[0024] Preferably, the bearing seat is substantially cylindrical and the rotor housing is substantially cylindrical, and a central axis of the bearing seat is offset from a central axis of the rotor housing.

[0025] It should be understood that "base wall thickness" refers to the thickness of the rotor housing wall in the area closest to the floor when in use, and "top wall thickness" refers to the thickness of the rotor housing wall in the area furthest from the floor when in use.

[0026] Preferably, the interior profile of the rotor housing is asymmetric.

[0027] It has been found that a larger wall thickness at the base of the rotor housing effectively lifts the wheel and compensates for the overhanging effect, i.e., the effect of gravity on the overhanging wheel, thereby reducing potential problems in the spinning process by adjusting the rotor to a desired position. For example, potential problems arise when various auxiliary equipment, such as air nozzles or binder delivery nozzles, are not aligned with the wheel.

[0028] Preferably, the clearance between the annular bearing seat and the inner surface of the rotor housing is from about 10 mm to about 18 mm, more preferably from about 12 mm to about 16 mm, and most preferably about 14 mm.

[0029] It has been found that by increasing the clearance between the annular bearing seat and the inner surface of the rotor housing, the risk of failure due to debris / slug lodged between the annular bearing seat and the inner surface of the rotor housing is significantly reduced. If debris / slug is lodged in the clearance, the suspension can no longer move and the bearings are damaged. The configuration of the present invention ensures that this cause of failure is eliminated.

[0030] Preferably, the height of each damper is about 20 mm to about 30 mm, more preferably, the height of each damper is about 22 mm to about 27 mm, and most preferably, the height of each damper is about 25 mm.

[0031] Preferably, the outer surface of each damper has a diameter of about 18 mm to about 22 mm, more preferably, the outer surface of each damper has a diameter of about 19 mm to about 21 mm, and most preferably, the outer surface of each damper has a diameter of about 20 mm.

[0032] It should be understood that the "external" surface of the damper refers to the surface adjacent the rotor housing.

[0033] Preferably, the inner surface of each damper has a diameter of about 25 mm to about 29 mm, more preferably, the inner surface of each damper has a diameter of about 26 mm to about 28 mm, and most preferably, the inner surface of each damper has a diameter of about 27 mm.

[0034] It should be understood that the "inner" surface of the damper refers to the surface adjacent the bearing seat.

[0035] Preferably, the total volume of each damper is about 35,000 mm3 to about 45,000 mm3, more preferably, the total volume of each damper is about 39,000 mm3 to about 44,000 mm3, and most preferably, the total volume of each damper is about 43,000 mm3.

[0036] Preferably the or each damper is a rubber damper.

[0037] Optionally, the or each damper is a silicone damper.

[0038] Preferably the or each damper is a neoprene rubber damper.

[0039] Preferably the or each damper has a Shore A hardness of between 40 and 60, more preferably the or each damper has a Shore A hardness of about 55.

[0040] Preferably, the damping stiffness is about 5·105 N / m to 106 N / m, and more preferably, the damping stiffness is 106 N / m or less.

[0041] "Dampering stiffness" is understood to be the total stiffness of the total number of dampers arranged in a ring.

[0042] It has been found that if the damping stiffness is too low, this will cause the wheel to sag to a greater extent than desired, while if the damping stiffness is too high, the bearing life will be shortened. Furthermore, if the damping stiffness is too low, the rotor will move more than desired, which can cause damage as there is significant movement to the coupling to the motor or to the contact between the parts. By optimizing the damping stiffness, the rotor vibration and imbalance can be accurately compensated, reducing wear to the bearings and increasing the rotor life. Rigorous testing has shown that a larger volume of softer rubber works more effectively than a smaller volume of stiffer rubber. The damping stiffness of the present invention has been optimized for a working rotational speed of about 4000 RPM to 13000 RPM. Bearing life is how long a user can expect a ball bearing to last under standard operating conditions, which has been found to depend on the amount of bearing load, and is calculated in revolutions, so that the time per revolution and the percentage of time the bearing is continuously rotating are used to determine the bearing life.

[0043] Preferably, the first bearing assembly is at the non-drive end of the rotor and comprises between 10 and 24 dampers, more preferably the first bearing assembly is at the non-drive end of the rotor and comprises 20 dampers. Preferably, the second bearing assembly is at the drive end of the rotor and comprises between 10 and 24 dampers, more preferably the second bearing assembly is at the drive end of the rotor and comprises 18 dampers.

[0044] The volume of rubber and number of dampers in the "soft" suspension of the present invention are carefully selected to withstand wear. For every rotor size, the optimum number of rubber dampers is used to provide the required life while ensuring that any imbalance is compensated for.

[0045] Preferably, the rotor includes between about 10 and about 24 annularly arranged frusto-conical dampers, and more preferably, the rotor includes between about 10 and about 24 annularly arranged frusto-conical dampers substantially equidistant from one another around the annular bearing assembly.

[0046] Preferably the or each bearing is a ball bearing, more preferably an angular contact ball bearing.

[0047] Preferably the or each bearing is a hybrid angular contact ball bearing having a steel lining and balls made from a ceramic material.

[0048] Preferably, the inner diameter of the or each ball bearing is from about 40 mm to about 80 mm, more preferably the diameter of the or each ball bearing is from about 60 mm to about 70 mm, and most preferably the diameter of the or each ball bearing is about 70 mm.

[0049] A smaller diameter increases the life of the bearings, but a diameter that is too small is problematic in terms of fitting the wheel onto the shaft (the contact surface on the shaft becomes too small).

[0050] Preferably, the bearing assembly comprises two spaced apart angular contact ball bearings.

[0051] Preferably, the distance between the two angular contact bearings is about 10 mm to about 30 mm, more preferably, the distance between the two angular contact bearings is about 15 mm to about 25 mm, and most preferably, the distance between the two angular contact bearings is about 20 mm.

[0052] Preferably, the contact angle of each angular contact ball bearing is approximately 15°.

[0053] Preferably, the bearing assembly comprises two angular contact ball bearings spaced apart by inner and outer axial spacer rings.

[0054] It has been found that when the rotor is in use, there is a significant temperature difference between the rotor shaft and the bearing seat. When cold, the shaft will have a smaller diameter and a larger pressure angle relative to the pressure direction against the bearing. When warm, the shaft will expand to a larger diameter and the pressure angle against the bearing will decrease. By configuring the bearing assembly with an inner axial spacer ring and an outer axial spacer ring, the expected temperature difference is allowed so that the ball bearings do not "rattle" or are subjected to much greater pressure, but are in the desired position.

[0055] Preferably, the width of the outer spacer ring is less than the width of the inner spacer ring.

[0056] Preferably, the width of the outer spacer ring is about 10 μm to about 70 μm less than the width of the inner spacer ring, and more preferably, the width of the outer spacer ring is about 61 μm less than the width of the inner spacer ring.

[0057] Preferably the or each spacer ring is steel.

[0058] Preferably, the shaft is substantially cylindrical.

[0059] Preferably, the outer cross-sectional diameter of the shaft is from about 80 mm to about 120 mm, more preferably about 100 mm.

[0060] The diameter of the shaft of the present invention is a compromise, since increasing the diameter makes the shaft stiffer, thereby positively affecting the dynamic behavior of the system, but negatively affecting weight and cost. If a diameter of 30 mm was chosen for the present system, the flexibility of the shaft would mean that it would rotate at a critical speed of 12,000 RPM and bend critically.

[0061] Preferably, the relationship between shaft diameter (Dshaft) and shaft length (Lshaft) is defined as Dshaft(Lshaft)≧0.12*Lshaft−32mm for a shaft length range of about 101 mm to about 1325 mm, and for a shaft diameter range of 20 mm or more, and for a seat stiffness (damping stiffness) of 3*106 N / m or less.

[0062] By increasing the cross-sectional diameter, also referred to as the "thickness" of the shaft, the vibrations generated by the shaft as it rotates are significantly reduced. Reducing the vibrations reduces wear on the moving parts and device imbalance, which in turn increases the mean time between failures.

[0063] Preferably, the length of the shaft between the centre point of the first bearing assembly and the centre point of the second bearing assembly is between about 530 mm and about 590 mm, more preferably about 590 mm.

[0064] Preferably, the overall length of the shaft is about 800 mm to about 1200 mm, preferably about 1000 mm.

[0065] Preferably the shaft is steel.

[0066] Preferably the or each bearing seat weighs from about 1.5kg to about 3.5kg, preferably from about 2kg to about 3kg, more preferably each bearing seat weighs about 3kg.

[0067] It has been found that reducing the mass of the bearing seat reduces vibration and resulting wear on the bearing, thereby increasing the bearing life and mean time between failures.

[0068] Preferably, the bearing seat is an annular ring having a plurality of substantially cylindrical recesses each for receiving a damper, preferably a frusto-conical damper. Optionally, the bearing seat is an annular ring having a plurality of truncated cylindrical recesses each for receiving a damper, preferably a frusto-conical damper.

[0069] By minimizing the weight of the bearing seat, the load on the bearing is reduced. The shape and configuration of the bearing seat securely retains the damper while allowing easy removal of the damper for maintenance and to access the ball bearings.

[0070] Preferably, the maximum rotational speed of the rotor is about 13,000 RPM.

[0071] Preferably, the rotational speed of the rotor is from about 6,000 RPM to about 13,000 RPM.

[0072] Preferably, the rotor further comprises a cooling system.

[0073] Preferably, the cooling system comprises at least one fluid inlet and at least one fluid outlet, with at least one channel therebetween passing through at least one bearing seat of the rotor.

[0074] The cooling system of the present invention allows the temperature change (ΔT) between the bearing seats to be substantially constant. Thus, the temperature of the ball bearings can be reduced. The cooling system also ensures that the temperature of the rubber damper is kept low, so that the maximum temperature is maintained at about 50-60°C.

[0075] Preferably the rotor is water cooled.

[0076] Preferably, the rotor further comprises an air flow system.

[0077] Preferably, the rotor further comprises an air flow purge system.

[0078] The present invention operates in a harsh environment and the bearings are relatively exposed due to the open housing design, and it has been found that air flow through the system can be used to remove unwanted debris and contaminants from around the or each bearing, reducing uneven bearing wear and optimizing rotor performance.

[0079] Preferably, each rotor is provided with a drive means.

[0080] Preferably, the rotor housing is substantially cylindrical.

[0081] More preferably, the rotor housing is substantially cylindrical comprising two mating parts. Preferably, the two mating parts are substantially symmetrical. More preferably, the rotor housing comprises two semi-cylindrical shells. Preferably, the two semi-cylindrical shells are mated with each other to form a substantially cylindrical housing. Preferably, each shell has the shape of a longitudinal half of a cylinder.

[0082] By providing an open housing that can be easily and conveniently opened, maintenance time and complexity is reduced, and as a result "downtime" during which the device is not operational for maintenance is also reduced.

[0083] In a further aspect, the present invention provides a yarn harvesting apparatus comprising a set of at least three rotors as described herein, each rotor mounted to rotate about a different substantially horizontal axis and arranged such that as the rotors rotate, melt cast onto the periphery of a first rotor in the set is successively cast onto the periphery of each subsequent rotor and fibre is discharged from the rotors.

[0084] Preferably, the thread harvesting device comprises a set of four rotors as described herein.

[0085] Preferably, each subsequent rotor is sized so that it is capable of providing greater acceleration than the preceding rotor in the set.

[0086] Preferably, each rotor is mounted on a wheel.

[0087] Preferably, the first rotor is attached to a first wheel having a diameter of about 184 mm, the first wheel being rotatable at about 5,000 RPM to about 6,000 RPM in an acceleration field of about 25 km / s 2 to about 36 km / s 2 .

[0088] Preferably, the second rotor is attached to a second wheel having a diameter of about 234 mm, the second wheel being rotatable at about 6,000 RPM to about 13,000 RPM in an acceleration field of about 46 km / s 2 to about 217 km / s 2 .

[0089] Preferably, the third rotor is attached to a third wheel having a diameter of about 314 mm, the third wheel being rotatable at about 6,000 RPM to about 13,000 RPM in an acceleration field of about 62 km / s 2 to about 291 km / s 2 .

[0090] Preferably, the fourth rotor is attached to a fourth wheel having a diameter of about 332 mm, the fourth wheel being rotatable at about 6,000 RPM to about 13,000 RPM in an acceleration field of about 65 km / s 2 to about 308 km / s 2 .

[0091] Preferably, the yarn harvesting device further comprises a collector, more preferably a chamber for collecting the fibres from the or each rotor and removing them from the set of rotors.

[0092] Preferably, the thread harvesting device further comprises at least one temperature sensor, and optionally a pyrometer.

[0093] In a further aspect, the present invention provides a method of producing man-made vitreous fibers (MMVF), comprising the steps of: providing a yarn harvesting apparatus comprising a set of at least three rotors as described herein, each mounted to rotate about a different substantially horizontal axis, each rotor having a drive means; rotating the rotors; providing a mineral melt to form man-made vitreous fibers (MMVF), the melt being cast onto the circumference of the first rotor; and collecting the formed fibers.

[0094] For clarity and conciseness of description, features may be described herein as part of the same or separate embodiments, however, it should be understood that the scope of the present invention may include embodiments having all or a partial combination of the described features.

[0095] The invention will now be described, by way of example only, with reference to the accompanying drawings, in which: [Brief description of the drawings]

[0096] [Figure 1] 1 is a longitudinal section through a rotor according to the invention; [Diagram 2] FIG. 2 is a perspective view of a right position spinner with four rotors according to the present invention, shown without the spinner housing. [Diagram 3] FIG. 2 is a perspective view of a bearing assembly at the non-drive end (NDE) of the rotor of FIG. [Figure 4a] FIG. 2 is a perspective view of a bearing seat of a bearing assembly of a rotor according to the invention, without showing the damper; [Figure 4b] FIG. 4b is a perspective view of a section of the bearing seat shown in FIG. 4a. [Figure 5a] FIG. 4 is a cross-sectional view through the frusto-conical damper shown in the bearing assembly of FIG. [Figure 5b] FIG. 5b is a perspective view of a cross section of the frusto-conical damper of FIG. 5a. [Figure 5c] FIG. 5B is a perspective view of the damper of FIGS. 5a and 5b. [Figure 6a] FIG. 1 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. [Figure 6b] FIG. 6b is a perspective view of a section of the bearing seat shown in FIG. 6a; [Figure 7] FIG. 13 is a cross-sectional view through the bearing assembly at the non-drive end (NDE) of the rotor showing the larger clearance between the bearing assembly and the rotor housing. [Figure 8a] FIG. 2 is a perspective view of the rotor of the present invention showing half of the rotor housing removed. [Figure 8b] FIG. 2 is a plan view of an NDE of the rotor housing, bearing seat, and damper of the present invention. [Figure 9] Figures 9a, 9b and 9c are schematic cross-sections (not to scale) through a pair of angular contact ball bearings in a rotor of the present invention illustrating the preferred reduction in bearing preload due to temperature differential, with Figure 9b showing the shaft when cold and Figure 9c showing the shaft when warm. [Figure 10] FIG. 2 is an external perspective view of the rotor of the present invention showing the cooling and air purge system. [Figure 11] FIG. 1 is a cross-sectional view of a DE bearing seat showing the grease labyrinth. [Figure 12] FIG. 1 is a front view of the right position spinner of the present invention showing all four rotors with attached wheels. [Figure 13] FIG. 13 is a rear view of a 5D surface contour plot of rotor 4 simulating life for the NDE bearing seat for various shaft diameters, seat masses, shaft lengths, and seat rubber stiffness (damping stiffness) at a maximum rotor speed of 13,000 RPM. [Figure 14] Top view of the 5D surface contour plot of Figure 13 showing the mass of the seat (Mseat kg), shaft diameter of the mid-section (Dshaft mm), shaft length between seat centers (Lshaft mm), and seat rubber stiffness (Kseat N / m). [Figure 15]5D and contour plots of rotor 4 simulating the NDE seat displacement (ΔxseatNDE) at maximum rotor speed of 13000 RPM. DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

[0097] Referring to FIG. 1, a longitudinal cross section of rotor 1 is shown with shaft 2 longitudinally disposed between drive end (DE) 3 and non-drive end (NDE) 4 of rotor 1 such that the shaft 2 is substantially horizontal. Shaft 2 is a hollow cylindrical steel shaft having an outer diameter of about 100 mm and a bearing seat diameter of about 70 mm. In an alternative embodiment, the shaft outer diameter is about 100 mm to about 120 mm and the bearing seat diameter is about 50 mm to about 100 mm. The length of shaft 2 between the center point of the bearing at DE 3 and the center point of the bearing at NDE 4 is about 530 mm to 590 mm. For the illustrated embodiment, the length of shaft 2 is about 590 mm. At NDE 4, shaft 2 "overhangs" beyond first bearing assembly 5, and the overall length of shaft 2 is about 955 mm. A second bearing assembly 6 is disposed at DE 3. The bearing assemblies 5, 6 for both NDE 4 and DE 3 are identical (except for minor details such as the cooling water supply to the wheels), however the bearing assembly 5 of NDE 4 will be described in more detail with respect to Figures 3 to 5. The first and second bearing assemblies 5, 6 are each located adjacent a "soft" suspension, with each first and second soft suspension working independently of each other. The NDE 4 may comprise multiple bearing assemblies 5, which will increase bearing life as the static load on each bearing assembly 5 is reduced.

[0098] Referring to FIG. 2, a perspective view of the four rotors of a right position spinner is shown without the spinner body, showing a first rotor 1a, a second rotor 1b, a third rotor 1c, and a fourth rotor 1d. The right position refers to the location of the fourth rotor 1d on the right side. 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 located on the left side. The first rotor 1a has a maximum speed of 6000 RPM and is connected to the first motor 30a. The second rotor 1b has a maximum speed range of 6000-13000 RPM and is connected to the second motor 30b. The third rotor 1c has a maximum speed range of 6000-13000 RPM and is connected to the third motor 30c. The fourth rotor 1d has a maximum speed range of 6000-13000 RPM and is connected to the fourth motor 30d. Wheels 13b, 13c, 13d are shown attached to the second, third and fourth rotors respectively. Each motor 30a, 30b, 30c, 30d and the respective motor shaft are fixed in the spinner body. The rotor shaft 2 shown in FIG. 1 is flexibly mounted in the rotor housing 12, which is also fixed. A flexible lamellar coupling 32 connects the shafts of the drive end of each motor and the drive end of each rotor, allowing some radial misalignment. However, the maximum allowable radial misalignment of this coupling is 1.3 mm. The coupling therefore sets a limit on the allowable displacement of the shaft 2 at the DE 3. Referring to FIG. 8a, there is an additional connection between the motors 30a, 30b, 30c, 30d and the rotors 1, 1b, 1c, 1d to power and cool the spinner.

[0099] With reference to Figures 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, as shown in Figures 4a and 4b, having a plurality of recesses 9a spaced equidistantly around its outer surface with damper seats 9b spaced equidistantly therebetween. With reference to Figure 4b, each damper seat 9b has a threaded hole 9c in which a damper 11 is secured. As shown in Figure 4b, the recesses 9a are truncated cylindrical recesses. The bearing seat 8 has a very low mass of 3kg with minimal material to support the components held by the bearing seat 8. The bearing seat 8 also includes an aluminium labyrinth ring 10 to reduce seat mass.

[0100] With reference to Figures 3 and 4a, each damper seat 9b supports one end of a frusto-conical damper 11 that protrudes from the bearing seat 9b and the bearing seat 8. As shown in Figures 5a, 5b and 5c, the damper 11 is frusto-conical with a threaded metal screw 11a at a first, smaller, internal end face or tap end. The internal end face has a diameter of about 20 mm. At a second, external end face or bore end there is a further metal insert 11b, which has a central threaded opening 11c that is partially inserted into the rubber damper 11. The external end face has a diameter of about 27 mm and the length of the damper 11 is about 25 mm. With reference to Figures 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 with a threaded bore 9c in the bearing seat 8. Similarly, the outer end face of the damper 11 includes a threaded opening 11c for engaging a threaded screw or bolt (not shown) for connecting the damper 11 to the rotor housing. The damper 11 is made of neoprene rubber having a Shore A hardness of about 55 and a damping stiffness of about 106 N / m. It should be understood that the embodiment shown is for a bearing assembly / suspension in the NDE 4, but the damping stiffness of about 106 N / m is substantially similar in the DE 3. In alternative embodiments, the damper is cylindrical, but in the preferred embodiment shown in FIG. 3, there are 20 frusto-conical dampers 11 spaced equidistantly from each other around the bearing seat 8 for the rotor 4. Each damper 11 does not contact adjacent dampers 11 in an annular arrangement. The embodiment shown in FIG. 3 is for a fourth rotor of a thread harvesting device having four rotors. In an alternative embodiment, as shown in FIG. 6a, the rotor has about 12 to about 22 frusto-conical dampers 11' equidistant from one another.

[0101] Figures 6a and 6b show an alternative embodiment of the first bearing assembly 5' and bearing seat 8' for the third and fourth rotors having twelve substantially cylindrical dampers 11 equidistantly spaced from each other around the bearing seat 8', each damper 11 having a Shore A hardness of about 55 and a damping stiffness of about 106 N / m. The dampers 11 are connected to the bearing seat 8' by a screw connection, with a threaded screw of the dampers 11 engaging a threaded hole in the bearing seat 8'. Similarly, the other end of the dampers 11 comprises a threaded opening for engaging a bolt for connecting the dampers 11 to the rotor housing. The bearing seat 8' further comprises a ball bearing inner steel ring 16' and an aluminium labyrinth ring 10' used to reduce the seat mass.

[0102] 1, 2 and 7, the damper 11 is disposed 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 at both the NDE 4 and the DE (not shown). In use, the rotor 1 will be mounted to high strength steel wheels at location 13 further along the shaft 2 from the first bearing assembly 5 at the NDE 4. Referring to FIG. 2, the wheels 13a, 13b, 13c, 13d are disposed at a first end of the shaft 2, and the opposite end is connected to the high speed motors 30a, 30b, 30c, 30d by flexible couplings. 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 about 15 mm. It should be understood that the outer wall of the wheel is a curved surface and the side wall is circular and substantially perpendicular to the length of the shaft 2. Referring to FIG. 2, typically the wheel diameter of the wheel 3 is about 314 mm and has a mass of about 50 kg. For the preferred embodiment of the fourth rotor, the wheel is about 47 mm from the centerline of the NDE bearing assembly 5 and rotates at about 6000 RPM to about 13000 RPM. The rotor housing 12 supports the bearings and covers the middle portion of the shaft 2. However, for ease of understanding the invention, the wheel is not shown in FIG. 1.

[0103] 2 and 8a, each rotor housing 12 comprises two symmetrical mating parts, each of which is a semi-cylindrical shell, so that when the two parts are mated together, a substantially cylindrical housing 12 is formed. Each half of the rotor housing 12 is held to the adjacent half of the housing by a screw 33. FIG. 8a also shows a screw 11d which holds each damper (not shown) to the rotor housing. The rotor housing 12 further comprises a bracket 35 which supports the tubes and wires. Rubber rings 34 at each end 4, 5 of the rotor housing 12 are used to mount the rotor 1 in the spinner.

[0104] Referring to FIG. 8b, in the preferred embodiment, the rotor housing 12 has a greater wall thickness of about 5 mm at the base of the housing 12 compared to the wall thickness at the top of the rotor housing 12, which effectively lifts the wheels to compensate for the overhanging effect. The interior profile of the rotor housing 12 is asymmetric. Referring to FIG. 8b, the damper 11' at the top of the NDE 4 is stretched by about 2.5 mm. The damper 11'' at the bottom of the NDE 4 is compressed by about 2.5 mm. The offset location of the central axis of the annular mounting ring / bearing seat 8 with respect to the central axis of the annular rotor housing 12 counters the effect of gravity on the suspension of the bearings 7, the shaft 2, and the wheels 13 within the rotor housing 12, so that the center of gravity of the rotor 1 is in the desired location.

[0105] 3 and 6b, the inner surface of each annular bearing seat 8 supports two hybrid ball bearings 7, which are permanently lubricated ultra-precision hybrid angular contact ball bearings 7 selected to achieve the desired speed and life before failure. These two bearings 7 in each of the DE 3 and NDE 4 are fitted closely together at each end of the rotor 1 with very little space between them.

[0106] Referring to Figures 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 imbalance at the wheels of the rotor 4 has been determined to be approximately 560g·cm and the bearing seats and damper suspension configuration have been optimized to compensate for this imbalance. The angular contact ball bearings 7 have a diameter of approximately 70mm, although in alternative embodiments the bearing diameter is either 60mm, 65mm, 70mm or 75mm.

[0107] In the embodiment of Figures 9a, 9b and 9c, the ball bearings 7 have a diameter of 70 mm. Each pair of ball bearings 7 is axially separated by an inner spacer ring 14 between the rotating inner surfaces of the ball bearings 7 adjacent 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 and a ball bearing outer steel ring 17 mounted on the steel shaft 2. Each angular contact ball bearing 7 has an angular contact (pressure) angle β that is symmetric about the centre line between the ball bearings 7. The contact angle β is in the pressure direction during rotation. The contact angle β will vary depending on the temperature of the steel shaft 2.

[0108] With reference to FIG. 9b, when the shaft 2 is colder, the contact angle β will increase because the shaft will contract to a smaller diameter. With reference to FIG. 9c, when the shaft 2 is warmer, the contact angle β will decrease because the shaft will expand to a larger diameter. As shown in FIG. 9b, when the shaft 2 is cold, it has a smaller diameter Rc and the pressure angle β is larger. The outer spacer ring 15 has a width Wc and does not contact either the ball bearing inner steel ring 16 or the ball bearing outer steel ring 17, and there is a larger range for the movement / rattle of the ball bearing 7. However, as shown in FIG. 9c, when the shaft 2 is warm, it has a larger diameter RH and the pressure angle β is smaller. The increased diameter of the shaft also means that the bearing balls 7 will push the outer bearing seat 17 towards each other until they come into contact with the outer spacer ring 15. Thus, the bearing assemblies 5, 6 are configured to reduce the bearing preload due to the expected temperature difference when the rotor is in use, which is about 10°C.

[0109] 9b and 9c, it has been found to be advantageous that the outer spacer ring 15 is shorter than the inner spacer ring 14. Typically, for a bearing diameter of 70 mm for rotors 3 and 4, the width of the outer spacer ring 15 is approximately 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 approximately 16 μm less than the inner spacer ring 14.

[0110] As shown in FIGS. 3 and 10, the annular bearing seat 8 is surrounded by an end ring 18 .

[0111] As shown in Figure 10, the rotor 1 has a cooling system for both the DE bearing assembly 6 and the NDE bearing assembly 5. The NDE cooling fluid inlet 20 and the NDE cooling fluid outlet 21 are connected by a channel through which the cooling fluid flows to remove heat from the NDE 4 of the rotor 1. The DE cooling fluid inlet 23 and the DE cooling fluid outlet 24 are also connected by a channel to remove heat from the DE 3 of the rotor 1. A fluid inlet 25 for water cooling the wheel (not shown) is also provided. In the illustrated embodiment, the cooling fluid is water. It has been shown through testing that the bearing temperatures can all be maintained at about 50°C.

[0112] 10, rotor 1 further includes DE air purge inlet 26 and NDE air purge inlet 27. The air purge system flushes contaminants from rotor 1. Rotor 1 further includes accelerometers 28 and 29 at NDE 4 and DE 3, and pyrometers 30 and 31 at DE 3 and NDE 4.

[0113] 10 and 11, an air purge system flushes contaminants from the rotor when the spinner is in use. The rotor also includes a water cooling system with a cooling water outlet 24 for removing heated water from the bearings 7. It has been found that when the spinner is washed during routine maintenance, water can contaminate the bearings 6. Therefore, the present invention further includes a grease labyrinth 38 to prevent water from contaminating the bearings 7 during washing of the spinner, as shown in FIG. 11. FIG. 11 shows a bearing assembly 6 of a DE4 with a rubber ring 34 for mounting the rotor 1 in the spinner. In the case of the DE3 shown, there is a DE labyrinth grease inlet 38a and an NDE labyrinth grease inlet 38b. The NDE labyrinth grease inlet 38b conveys grease to the NDE4, which has a similar grease labyrinth. When the rotor is stationary because the spinner is being maintained, the labyrinth 38 is sealed with grease and prevents water from contaminating the bearings 7 during washing. Sealing ring 39 prevents labyrinth grease from flowing into bearings 7 which are permanently lubricated with their own grease.

[0114] Referring to FIG. 12, in use, a yarn harvesting machine comprising four rotors 1a, 1b, 1c, 1d is used to produce man-made vitreous fibres (MMVF). Each rotor 1a, 1b, 1c, 1d is mounted to rotate about a different substantially horizontal axis, and each rotor 1a, 1b, 1c, 1d has a drive means, which may be a single drive means for powering all four rotors 1. In a preferred embodiment, the second rotor comprises a suspension ring bearing assembly at the non-drive end having 14 neoprene rubber dampers, and the third and fourth rotors at the non-drive end have bearing assemblies having 20 neoprene rubber dampers. The second rotor comprises 12 neoprene rubber dampers in the drive end bearing assembly, and the third and fourth rotors have 18 neoprene rubber dampers in the drive end bearing assembly.

[0115] Each right position rotor 1a, 1b, 1c, 1d has a high-speed motor connected at one end by a flexible coupling, and wheels are arranged at the opposite end of the rotors 1a, 1b, 1c, 1d. In the embodiment shown in FIG. 12, rotors 1 and 3 rotate counterclockwise, and rotors 2 and 4 rotate clockwise. The wheels of the first rotor 1a have a diameter of about 184 mm, the wheels of the second rotor 1b have a diameter of about 234 mm, the wheels of the third rotor 1c have a diameter of about 314 mm, and the wheels of the fourth rotor 1d have a diameter of about 332 mm. The space between the outer surfaces of each mounted wheel is varied, with the largest distance of about 228 mm between the wheels of the first and fourth rotors 1a, 1d, and the smallest distance of about 17 mm between the wheels of the first and second rotors 1a, 1b.

[0116] Referring to Figure 12, as the rotors 1a, 1b, 1c, 1d rotate, molten mineral melt of stone or rock or slag or glass melt is poured through the inlet 36 onto the circumference of the wheel of the first rotor 1a, spinning the melt and discharging the MMVF. The melt is then successively fed onto the wheels of the remaining rotors 1b, 1c, 1d, forming and collecting fibers each time. At the same time, as the melt is passed to the wheels of each successive rotor 1a, 1b, 1c, 1d, a high pressure air flow through the spinner and along the wheel removes the fibers from the wheel for collection.

[0117] For a typical four rotor yarn harvester according to the 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%. Production using the rotors of the invention can continue for about 4000 hours before the ball bearings 7 require replacement, which is a significant increase over known devices. Endurance tests comparing a prior art spinner operating at maximum speed (9300 RPM) with a spinner according to the invention operating at 13000 RPM, both operating with an unbalance of 560 g·cm, showed an improvement from 603 hours to over 4000 hours. Further tests found a mean time between failures of about 15000 hours for the spinner according to the invention.

[0118] 13, to evaluate the optimal parameters for the improved rotor of the present invention, a range of shaft lengths (Lshaft), shaft diameters (Dshaft), damping stiffness (Kseat), bearing seat mass (Mseat), and bearing diameters were simulated. The effect of these values ​​on bearing life (L10) and rotor shaft displacement (Δx) is shown in a 5D plot.

[0119] Figure 13 is a rear view of a 5D contour plot for rotor 4 at maximum rotor speed of 13000 RPM for an NDE seat with a bearing diameter of 70 mm. The height of each peak (L10) represents the bearing life in hours or the Basic Life Rating. The bearing life statistic used is a measure of the amount of time in rotation when 90% of the ball bearings can be expected to remain.

[0120] As shown in Figure 13, the optimal configuration of the present invention allows for a maximized bearing life of 3617 hours. The simulation considers a range of shaft diameters (Dshaft) and bearing seat mass (Mseat) plotted along with the shaft length between bearing seat centers (Lshaft) and the rubber stiffness at the bearing seat (Kseat). Each cube shown in the 5D plot of Figure 13 represents a 3D plot of life (L10) as a function of shaft length (Lshaft) and shaft diameter (Dshaft). Each 3D subplot has a rubber stiffness (Kseat) and bearing seat mass (Mseat) constant indicated by the location of the 3D subplot.

[0121] Figure 14 shows a top view of the 5D reface and contour plots of the rotor 4 shown in Figure 13. Each square in the plot represents a shaft length range of 101 to 1325mm in increments of 101mm, 407mm, 713mm, 1019mm, and 1325mm, and a shaft diameter range of 20 to 170mm in increments of 20mm, 58mm, 95mm, 133mm, and 170mm. The rubber stiffness (Kseat) ranges from 104N / m to 108N / m, and the bearing seat mass (Mseat) ranges from 1.5 to 12.0kg.

[0122] 13 and 14, the rubber stiffness (Kseat) at the bearing seat is optimized at about 106 N / m or less. Also, the bearing seat mass (Mseat) is preferably as low as possible, and in the case of the reduced mass of the bearing seat of the present invention of 3 kg, the bearing life is improved. Also, it has been found that if the seat mass (Mseat) is significantly reduced, for example to about 1.5 kg, it will be necessary to reduce the rubber stiffness (Kseat) at the bearing seat in order to achieve an increase in the bearing life. Also, it has been shown that an improvement in the bearing life is achieved for the present invention in the case of a shaft length (Lshaft) of 590 mm and a shaft diameter (Dshaft) of 100 mm. A further embodiment of the present invention with a bearing diameter of 60 mm has been investigated, and it has been found that the present invention can achieve a maximum bearing life of 6183 hours.

[0123] Referring to FIG. 15, a 5D surface contour plot for rotor 4 is shown plotting the NDE seat displacement (ΔxseatNDE) at a maximum rotor speed of 13000 RPM for an NDE seat with a bearing diameter of 70 mm. As shown, for a maximum allowable displacement (ΔxseatNDEmax) of 15 mm, the seat rubber stiffness is 105 N / m. Therefore, it was concluded that the seat rubber stiffness (damping stiffness) should be greater than 105 N / m to avoid exceeding the maximum allowable bearing seat displacement.

[0124] As used herein, the term "about" means plus or minus 20%, more preferably plus or minus 10%, even more preferably plus or minus 5%, and most preferably plus or minus 2%.

[0125] As used herein, the term "substantially" means a deviation of plus or minus 20%, more preferably plus or minus 10%, even more preferably plus or minus 5%, and most preferably plus or minus 2%.

[0126] The above embodiments are given by way of example only and those skilled in the art will of course appreciate that many modifications can be made without departing from the scope of the claims.

Claims

1. A rotor for a thread collection device, comprising: A rotor housing; first and second bearing assemblies, each bearing assembly including 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; a rotor comprising a plurality of resilient dampers disposed in an annular ring, each resilient damper connected to and engaging the bearing seat at a first end and connected to and engaging an inner wall of the rotor housing at a second end.

2. A rotor according to claim 1 , wherein each elastic damper is a frustum, preferably each elastic damper is a truncated cone.

3. 3. A rotor as claimed in claim 1 or 2, comprising a plurality of frusto-conical resilient dampers, each having a larger diameter at a face of the damper adjacent an inner wall of the rotor housing and a smaller diameter at a face of the damper adjacent the bearing seat, and / or the or each damper is a rubber damper, a silicone damper or a neoprene rubber damper.

4. A rotor according to any one of claims 1 to 3, wherein the plurality of elastic dampers are adapted to a working rotational speed of the rotor of between about 4000 RPM and 13000 RPM.

5. 5. A rotor as claimed in any one of claims 1 to 4, wherein the rotor housing has a greater wall thickness at its base than at its top surface, preferably the rotor housing has a base wall thickness of about 5mm greater than the wall thickness at its top surface, and / or the internal profile of the rotor housing is asymmetric, and / or the bearing seat is substantially cylindrical, the rotor housing being substantially cylindrical and a central axis of the bearing seat is offset from a central axis of the rotor housing.

6. 6. A rotor as claimed in any preceding claim, wherein the or each damper has a Shore A hardness of about 55 and / or the damping stiffness is between about 5·105 and about 106 N / m.

7. 7. A rotor as claimed in any one of claims 1 to 6, wherein each damper comprises a threaded screw for releasable connection to the bearing seat, and / or each damper comprises a threaded opening through the rotor housing for releasable connection with a screw.

8. A rotor as claimed in any preceding claim, comprising between about 10 and about 24 frusto-conical dampers arranged in an annular arrangement, preferably said dampers being substantially equidistant from one another around the annular bearing assembly.

9. 9. A rotor as claimed in any one of the preceding claims, wherein the or each bearing is a hybrid angular contact ball bearing, and / or the inner diameter of the or each ball bearing is between about 60 mm and about 75 mm, and / or the bearing is made of a ceramic material, and / or the distance between two angular contact bearings is about 20 mm.

10. The clearance between the annular bearing seat and the inner surface of the rotor housing is about 14 10. A rotor according to claim 1, wherein the angular contact ball bearings are spaced apart by an inner axial spacer ring and an outer axial spacer ring, preferably the width of the outer spacer ring is smaller than the width of the inner spacer ring, more preferably the width of the outer spacer ring is approximately 16 μm to 61 μm smaller than the width of the inner spacer ring.

11. A rotor according to any one of claims 1 to 10, wherein the relationship between shaft diameter (Dshaft) and shaft length (Lshaft) is defined as Dshaft(Lshaft) ≥ 0.12*Lshaft - 32mm for a shaft length range of approximately 101 mm to 1325 mm, and for a shaft diameter range of 20 mm or more, and for a seat stiffness (damping stiffness) of 3*106 N / m or less.

12. 12. A rotor as claimed in any one of claims 1 to 11, wherein the outer cross-sectional diameter of the shaft is approximately 100 mm, and / or the shaft has a bearing seat diameter of approximately 70 mm, and / or the length of the shaft is approximately 955 mm, and / or the length of the shaft between the centre points of the first and second bearing assemblies is approximately 590 mm.

13. A rotor as claimed in any preceding claim, wherein the or each bearing seat weighs no more than 3kg.

14. 14. A yarn harvesting machine comprising a set of at least three rotors according to any one of claims 1 to 13, each rotor mounted to rotate about a different substantially horizontal axis and arranged such that, as the rotors rotate, melt flowing onto the circumference of a first rotor in the set is successively deposited onto the circumference of each of the succeeding rotors and fibre is discharged from said rotors.

15. 1. A method for producing man-made vitreous fiber (MMVF), comprising the steps of: - providing a thread harvesting device comprising a set of at least three rotors according to any one of claims 1 to 13, each mounted for rotation about a different substantially horizontal axis, each rotor having a drive means; rotating the rotor; providing a mineral melt to form a man-made vitreous fiber (MMVF), the melt being cast onto the periphery of the first rotor; and collecting the formed fibers.

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