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, and more particularly to a rotor for a take-up device for use in the production of man-made vitreous fibers (MMVF), and to a method for the production of man-made vitreous fibers (MMVF).

Background Art

[0002] A spinning device known as a take-up device or (cascade) spinner is used in the production of MMVF for producing insulating materials from mineral melts or slag melts or glass melts of, for example, stone or rock, or for providing sound insulating or heat insulating materials. The take-up device has a set of rotors for spinning molten material or lava on a spinning wheel to produce a web-shaped insulating product. Molten stone or lava ("melt") is continuously fed from a first rotor to the remaining rotors of the set, and fibers are discharged from each wheel as each rotor rotates. These fibers are collected for the production of insulating products such as stone wool insulating products and removed from the set of rotors.

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

[0004] Known spinners operate at high speeds and accelerations of approximately 150 km / s² to achieve the very fine fibers required for good thermal insulation properties. Each rotor wheel has a rotating shaft suspended between bearings at the drive end (DE) and non-drive end (NDE), respectively. The NDE and DE of the rotor are not equidistant from the respective ends of the shaft, as the shaft at the NDE passes through the wheel into which the molten material is guided beyond the rotor. 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 on the bearings seated at each end of the rotor, and on dampers positioned between the rotor housing and the spinner body. Known devices use dampers positioned between the rotor housing and the spinner body to reduce vibrations transmitted from one rotor to the other in a set.

[0005] The use of springs as a means of vibration absorption is known. U.S. Patent No. 2,556,317 discloses a bearing assembly for a centrifuge in which a radial compression spring or rubber cushion is radially positioned between the bearing element and the stationary frame of the machine. WO2014 / 000799 discloses a spring damping element for an electric compressor / turbine generator. The damping element is a spring steel ring having a recess from which a leaf spring is received to apply a radial force.

[0006] However, the spinning scale and speed of the yarn-reeling equipment place very high loads on the bearings used, resulting in the need for frequent replacement of worn bearings. Typically, a four-wheel spinner using the rotor of the present invention produces 5-6 tons of stone wool per hour, and as a result, the reduction in "downtime" due to maintenance significantly increases the volume of product that can be produced.

[0007] The rotors within a spinner are each positioned around a substantially horizontal axis, resulting in increased wear on the rotor bearings due to gravity contributing to the imbalance, which leads to variable wear on the bearings. The imbalance in the forces applied to each rotor is also caused by the melt being fed onto the rotor wheel. Uneven wear of the rotor mechanics, or wear on the outer surface of the rotor assembly through which the molten material is guided, exacerbates the imbalance. One example is the accumulation of a layer of solidified melt on the rotor, so-called "freeze lining," which can be uneven and can detach in areas causing the imbalance. The combination of these factors causes the bearings used in existing spinners to wear out much faster than desired, necessitating the removal of the spinner from production for maintenance. Therefore, there is a strong need to improve the rotor configuration and dynamics to increase the mean time between failures. [Overview of the Initiative] [Problems that the invention aims to solve]

[0008] The present invention aims to provide an improved rotor for a silk reeling device that addresses the above-mentioned problems related to imbalance and high rotational force that cause wear on the rotor bearings. [Means for solving the problem]

[0009] In a first aspect, the present invention relates to a rotor for a silk reeling device, comprising a rotor housing, first and second bearing assemblies, each bearing assembly having at least two ball bearings seated in their respective bearing seats, and a substantially horizontal shaft rotatably mounted between the first and second bearing assemblies, wherein a plurality of elastic dampers are annular Lined up The present invention provides a rotor in which each elastic damper is connected to and engaged with the bearing seat at a first end and connected to and engaged with the inner wall of the rotor housing at a second end.

[0010] In the context of this invention, it should be understood that “rotor” refers to a rotating assembly, “shaft” refers to a long cylindrical rotating rod used to transmit rotational power, “damper” refers to a device for suppressing or absorbing vibrations, and “internal wall” of the rotor housing refers to the wall facing the rotor shaft.

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

[0012] The present invention reduces wear on the bearings, and as a result, the rotor can withstand the high speed and significant loads applied by the spinner. The annular arrangement of multiple elastic dampers is particularly well suited to the instability that occurs 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 adequately protect the bearings 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 bearing / rotor, and as a result, improves the efficiency of the yarn-throwing device by reducing maintenance time. By releasably connecting the elastic dampers to the bearing seats and the inner walls of the rotor housing, the dampers are made to work in both compression and tension, resulting in a significant increase in the mean time between failures of the bearings. Potential problems associated with the natural frequencies of the dampers of the present invention are avoided, and loss of action that can occur when using springs, for example, if the spring loses contact at one end, is avoided.

[0013] It has been found that by using multiple frustoconical dampers, the load on the bearings can be significantly reduced, even when the rotor speed reaches the critical overspeed for the bearings. For example, tests at 13,500 RPM with and without the soft suspension of the present invention showed a dynamic load reduction of 2,760 N to 192 N.

[0014] Multiple frustoconical dampers minimize internal bearing wear by providing improved absorption of vibrations generated by the high speed / high acceleration and unbalance of the spinning rotor. In effect, each damper is positioned and shaped to provide more elastic material where most is required to withstand both static and dynamic loads applied to the bearing assembly. It has been demonstrated that frustoconical, conical, or frustoconical dampers better withstand both static and dynamic loads in bearing suspensions. Unbalances that occur when the shaft is rotating at high speed, for example, when melt is being poured in or fibers are being discharged, are absorbed by elastic dampers that stretch or contract in accordance with the forces applied to the bearing assembly. The solutions of the present invention are particularly well-suited for use with spinning equipment and are carefully configured to optimize the volume and stiffness of the rubber placed in the annular ring to extend bearing life. The optimal choice has been found to maximize the volume of rubber 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 pyramidal pyramid. However, preferably, each elastic damper is rotationally symmetric.

[0016] The ease of mounting dampers in the spinner bearing seats is improved by the fact that the dampers have rotational symmetry, that is, that the dampers are rotationally symmetrical about their central axes.

[0017] Preferably, the rotor comprises a plurality of frustoconical elastic dampers that form an annular ring between the bearing seat and the inner wall of the rotor housing.

[0018] Preferably, the rotor comprises a plurality of frustoconical elastic dampers, each damper having a larger diameter at the inner wall of the rotor housing and a smaller diameter at the bearing seat.

[0019] Preferably, each damper is provided with a threaded screw for releasably connecting to a bearing seat, and / or each damper is provided with a threaded opening for releasably connecting to the screw through the rotor housing.

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

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

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

[0023] Preferably, the rotor housing has an increased base wall thickness and a decreased upper wall thickness, wherein the base wall thickness is increased by about 2 mm to about 3 mm and the upper wall thickness of the rotor housing is decreased accordingly. More preferably, the rotor housing has a base wall thickness increased by about 2.2 mm to about 2.7 mm and the upper wall thickness of the rotor housing is decreased accordingly. Most preferably, the rotor housing has a base wall thickness increased by about 2.5 mm and an upper wall thickness of the rotor housing that is decreased by about 2.5 mm compared to the standard wall thickness of the rotor housing.

[0024] Preferably, 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.

[0025] Please understand that "base wall thickness" refers to the thickness of the rotor housing wall in the area closest to the floor during use. "Upper wall thickness" refers to the thickness of the rotor housing wall in the area furthest from the floor during use.

[0026] Preferably, the internal profile of the rotor housing is asymmetrical.

[0027] A greater wall thickness at the base of the rotor housing has been found to effectively lift the wheel, compensating for the overhang effect, i.e., the action of gravity on the overhanging wheel, and as a result, reducing potential problems in the spinning process by adjusting the rotor to the desired position. For example, potential problems occur when various auxiliary facilities, such as air nozzles or binder supply 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 chips / slag remaining between the annular bearing seat and the inner surface of the rotor housing is significantly reduced. If chips / slag remain in the clearance, the suspension can no longer move and the bearing is damaged. The configuration of the present invention ensures that this cause of failure is eliminated.

[0030] Preferably, the height of each damper is from about 20 mm to about 30 mm, more preferably the height of each damper is from 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 from about 18 mm to about 22 mm, more preferably the outer surface of each damper has a diameter of from 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 "outer" surface of the damper refers to the surface adjacent to 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] Please understand that the "internal" surface of the damper refers to the surface adjacent to 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 damper or each damper is a rubber damper.

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

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

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

[0040] Preferably, the damping stiffness is about 5·10⁵N / m to 10⁶N / m, and more preferably, the damping stiffness is 10⁶N / 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, resulting in significant movement against the coupling or contact between components to the motor, which can cause damage. By optimizing the damping stiffness, rotor vibration and unbalance can be accurately compensated for, reducing wear on the bearing and increasing rotor life. Rigorous testing has shown that softer rubber in a larger volume works more effectively than stiffer rubber in a smaller volume. The damping stiffness of this invention is optimized for working rotational speeds of approximately 4000 RPM to 13000 RPM. Bearing life is how long a user can expect a ball bearing to last under standard operating conditions, and it has been found that this depends on the amount of bearing load, is calculated in rotational speed, and as a result, the time per revolution and the percentage of time the bearing is continuously rotating are used to determine bearing life.

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

[0044] The volume of rubber and the number of dampers in the "soft" suspension of this invention are carefully selected to withstand wear. For all rotor sizes, an optimal number of rubber dampers are used to provide the required lifespan while ensuring that any imbalance is compensated for.

[0045] Preferably, the rotor comprises about 10 to about 24 frustoconical dampers arranged in an annular manner. More preferably, the rotor comprises about 10 to about 24 frustoconical dampers arranged in an annular manner at equidistant from each other around a substantially annular bearing assembly.

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

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

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

[0049] Smaller diameters increase bearing life, but diameters that are too small pose a problem when considering the mounting of the wheel on 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 about 15 mm to about 25 mm, and most preferably about 20 mm.

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

[0053] Preferably, the bearing assembly comprises two angular contact ball bearings, each separated by an internal axial spacer ring and an external axial spacer ring.

[0054] It has been found that when a 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 with respect to the direction of pressure on the bearing. When warm, the shaft will elongate to a larger diameter, and the pressure angle on the bearing will decrease. By configuring the bearing assembly to include an internal axial spacer ring and an external axial spacer ring, the expected temperature difference is taken advantage of, and as a result, the ball bearing will not "rattle" or will be subjected to much greater pressure, but will remain in the desired position.

[0055] Preferably, the width of the external spacer ring is smaller than the width of the internal spacer ring.

[0056] Preferably, the width of the external spacer ring is about 10 μm to about 70 μm smaller than the width of the internal spacer ring, and more preferably, the width of the external spacer ring is about 61 μm smaller than the width of the internal spacer ring.

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

[0058] Preferably, the shaft is substantially cylindrical.

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

[0060] Increasing the diameter makes the shaft more rigid, which positively impacts the system's dynamic behavior, but negatively impacts weight and cost; therefore, the shaft diameter of this invention is a compromise. If a diameter of 30 mm is selected for this system, the shaft's flexibility means that it will bend decisively when rotating at a critical speed of 12,000 RPM.

[0061] Preferably, the relationship between the shaft diameter (Dshaft) and the shaft length (Lshaft) is defined as Dshaft(Lshaft) ≥ 0.12 * Lshaft - 32 mm for shaft lengths ranging from approximately 101 mm to approximately 1325 mm, shaft diameters of 20 mm or more, and seat stiffness (dampering stiffness) of 3 * 10⁶ N / m or less.

[0062] By increasing the cross-sectional diameter, also known as the "thickness" of the shaft, vibrations generated by the shaft during rotation are significantly reduced. Reducing vibrations reduces wear on moving parts and device imbalance, resulting in an increase in the mean time between failures.

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

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

[0065] Preferably, the shaft is made of steel.

[0066] Preferably, the weight of the bearing seat or each bearing seat is about 1.5 kg to about 3.5 kg, preferably about 2 kg to about 3 kg, and more preferably about 3 kg for each bearing seat.

[0067] It has been found that reducing the mass of the bearing seat reduces vibration and consequently wear on the bearing, which in turn increases the bearing's lifespan and mean time between failures.

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

[0069] Minimizing the weight of the bearing seat reduces the load on the bearing. The shape and configuration of the bearing seat securely holds the damper while allowing for easy removal of the damper for maintenance and access to the ball bearing.

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

[0071] Preferably, the rotor rotation speed is approximately 6,000 RPM to approximately 13,000 RPM.

[0072] Preferably, the rotor further includes 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 between them passing through at least one bearing seat of the rotor.

[0074] The cooling system of the present invention makes it possible to keep the temperature change (ΔT) between bearing seats substantially constant. Therefore, the temperature of the ball bearing can be reduced. In addition, the cooling system ensures that the temperature of the rubber damper is kept low, and as a result, the maximum temperature is maintained at about 50-60°C.

[0075] Preferably, the rotor is water-cooled.

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

[0077] Preferably, the rotor further comprises an airflow purging system.

[0078] This invention operates in harsh environments, and the bearings are relatively exposed due to the open housing design. The airflow through the system is used to remove undesirable debris and contaminants from or around the bearings, which has been found to reduce uneven bearing wear and optimize rotor performance.

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

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

[0081] More preferably, the rotor housing is substantially cylindrical and comprises two combined components. Preferably, the two combined components are substantially symmetrical. More preferably, the rotor housing comprises two semi-cylindrical shells. Preferably, the two semi-cylindrical shells combine 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 opened easily and conveniently, maintenance time and complexity are reduced, and as a result, "downtime" during which the device is not operational for maintenance is also reduced.

[0083] In a further embodiment, the present invention provides a yarn picking device comprising a set of at least three rotors as described herein, each rotor mounted to rotate around a different substantially horizontal axis, and arranged such that, as the rotors rotate, melt poured onto the circumferential portion of a first rotor in the set is continuously fed onto the circumferential portions of each subsequent rotor, and fibers are discharged from the rotors.

[0084] Preferably, the silk-collecting device comprises a set of four rotors as described herein.

[0085] Preferably, each subsequent rotor is sized to provide greater acceleration than the preceding rotor in the set.

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

[0087] Preferably, the first rotor is mounted on a first wheel having a diameter of approximately 184 mm, and the first wheel is rotatable at approximately 5,000 RPM to approximately 6,000 RPM in an acceleration field of approximately 25 km / s² to approximately 36 km / s².

[0088] Preferably, the second rotor is mounted on a second wheel having a diameter of approximately 234 mm, and the second wheel is rotatable at approximately 6,000 RPM to approximately 13,000 RPM in an acceleration field of approximately 46 km / s² to approximately 217 km / s².

[0089] Preferably, the third rotor is mounted on a third wheel having a diameter of approximately 314 mm, and the third wheel is rotatable at approximately 6,000 RPM to approximately 13,000 RPM in an acceleration field of approximately 62 km / s² to approximately 291 km / s².

[0090] Preferably, the fourth rotor is mounted on a fourth wheel having a diameter of approximately 332 mm, and the fourth wheel is rotatable at approximately 6,000 RPM to approximately 13,000 RPM in an acceleration field of approximately 65 km / s² to approximately 308 km / s².

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

[0092] Preferably, the silk-collecting device further comprises at least one temperature sensor and optionally a thermometer.

[0093] In a further embodiment, the present invention provides a method for producing artificial glass fibers (MMVF), comprising the steps of: providing a fiber picking apparatus comprising a set of at least three rotors described herein, each mounted to rotate around a different substantially horizontal axis, wherein each rotor has a driving means; rotating the rotors; providing a mineral melt to form artificial glass fibers (MMVF), the melt being poured onto the periphery of a first rotor; and collecting the formed fibers.

[0094] For clarity and conciseness, features are described herein as part of the same or separate embodiments. However, it should be understood that the scope of the invention may include embodiments having all or some combinations of the described features.

[0095] Next, the present invention will be described with reference to the attached drawings as an example. [Brief explanation of the drawing]

[0096] [Figure 1] This is a longitudinal cross-sectional view of the rotor according to the present invention. [Figure 2] A perspective view of a right-position spinner with four rotors according to the present invention, shown without the spinner housing. [Figure 3] Figure 1 is a perspective view of the bearing assembly at the non-drive end (NDE) of the rotor. [Figure 4a] This is a perspective view of the bearing seat of a rotor bearing assembly according to the present invention, with the damper not shown. [Figure 4b] Figure 4a is a perspective view of one section of the bearing seat. [Figure 5a] This is a cross-sectional view through the frustoconical damper shown in the bearing assembly in Figure 3. [Figure 5b] Figure 5a is a perspective view of the cross-section of the frustoconical damper. [Figure 5c] Figures 5a and 5b are perspective views of the damper. [Figure 6a] This is a perspective view of an alternative embodiment of the bearing assembly at the non-drive end (NDE) of the rotor according to the present invention. [Figure 6b] Figure 6a is a perspective view of one section of the bearing seat. [Figure 7] This is a cross-sectional view through the bearing assembly at the non-drive end (NDE) of the rotor, showing a larger clearance between the bearing assembly and the rotor housing. [Figure 8a] This is a perspective view of the rotor of the present invention showing half of the rotor housing after it has been removed. [Figure 8b] This is a plan view of the NDE of the rotor housing, bearing seat, and damper according to the present invention. [Figure 9] Figures 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, illustrating a favorable reduction in bearing preload due to temperature difference, with Figure 9b showing the shaft when cold and Figure 9c showing the shaft when warm. [Figure 10] This is an external perspective view of the rotor of the present invention, showing the cooling and air purging system. [Figure 11] This is a cross-sectional view of the DE bearing seat showing the grease labyrinth. [Figure 12] This is a front view of the right-position spinner of the present invention, showing all four rotors with mounted wheels. [Figure 13] This is a rear view of a 5D surface contour plot of rotor 4 simulating the lifespan of an NDE bearing seat for various shaft diameters, seat masses, shaft lengths, and seat rubber stiffness (dampering stiffness) at a maximum rotor speed of 13,000 RPM. [Figure 14] Figure 13 is a top view of the 5D surface contour plot, showing the seat mass (Mseat kg), shaft diameter in the middle section (Dshaft mm), shaft length between seat centers (Lshaft mm), and seat rubber stiffness (Kseat N / m). [Figure 15]This is a 5D and contour plot of rotor 4 simulating the displacement of the NDE seat (ΔxseatNDE) at a maximum rotor speed of 13000 RPM. [Modes for carrying out the invention]

[0097] Referring to Figure 1, a longitudinal cross-section of the rotor 1 is shown, in which the shaft 2 is longitudinally positioned between the drive end (DE) 3 and the non-drive end (NDE) 4 of the rotor 1 so as to be substantially horizontal. The shaft 2 is a hollow cylindrical steel shaft with an outer diameter of approximately 100 mm and a bearing seat diameter of approximately 70 mm. In alternative embodiments, the shaft outer diameter is approximately 100 mm to 120 mm, and the bearing seat diameter is approximately 50 mm to 100 mm. The length of the shaft 2 between the bearing center point at DE3 and the bearing center point at NDE4 is approximately 530 mm to 590 mm. In the embodiment shown, the length of the shaft 2 is approximately 590 mm. At NDE4, the shaft 2 "overhangs" beyond the first bearing assembly 5, and the total length of the shaft 2 is approximately 955 mm. The second bearing assembly 6 is positioned at DE3. The bearing assemblies 5 and 6 for both NDE4 and DE3 are identical (except for minor details such as the coolant supply to the wheels), but the bearing assembly 5 for NDE4 will be described in more detail with respect to Figures 3 through 5. The first and second bearing assemblies 5 and 6 are each positioned adjacent to a "soft" suspension, and the first and second soft suspensions operate independently of each other. NDE4 may have multiple bearing assemblies 5, which increases bearing life as the static load on each bearing assembly 5 is reduced.

[0098] Referring to Figure 2, a perspective view of the four rotors of a right-position spinner is shown without the spinner body, showing the first rotor 1a, the second rotor 1b, the third rotor 1c, and the fourth rotor 1d. Right position refers to the position 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 Figure 2, but with the fourth rotor 1d positioned 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 to 13000 RPM and is connected to the second motor 30b. The third rotor 1c has a maximum speed range of 6000 to 13000 RPM and is connected to the third motor 30c. The fourth rotor 1d has a maximum speed range of 6000 to 13000 RPM and is connected to the fourth motor 30d. Wheels 13b, 13c, and 13d are shown, attached to the second, third, and fourth rotors, respectively. Each motor 30a, 30b, 30c, and 30d, and their respective motor shafts, are fixed in the spinner body. The rotor shaft 2 shown in Figure 1 is flexibly mounted within the rotor housing 12, which is also fixed. A flexible lamellar coupling 32 connects the drive ends of each motor and the drive ends of each rotor shafts, allowing for some radial displacement. However, the maximum allowable radial displacement of this coupling is 1.3 mm. Therefore, the coupling limits the allowable displacement of shaft 2 in DE3. Referring to Figure 8a, there are further connections between the motors 30a, 30b, 30c, and 30d and the rotors 1, 1b, 1c, and 1d to power and cool the spinner.

[0099] Referring 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 shown in Figures 4a and 4b, having a plurality of recesses 9a equidistant around its outer surface, with damper seats 9b equidistant between them. Referring to Figure 4b, each damper seat 9b has a threaded hole 9c into which a damper 11 is fixed. As shown in Figure 4b, the recesses 9a are truncated cylindrical recesses. The bearing seat 8 has a very low mass of 3 kg, using minimal material to support the components held by the bearing seat 8. The bearing seat 8 also includes an aluminum labyrinth ring 10 to reduce the seat mass.

[0100] Referring to Figures 3 and 4a, each damper seat 9b supports one end of a frustoconical 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 frustoconical with a threaded metal screw 11a on the first smaller inner end face or tapped end. The inner end face has a diameter of approximately 20 mm. There is a further metal insert 11b on the second outer end face or bore end, which has a central threaded opening 11c that partially enters the rubber damper 11. The outer end face has a diameter of approximately 27 mm, and the length of the damper 11 is approximately 25 mm. Referring to Figures 3 and 4b, the damper 11 is connected to the bearing seat 8 by a screw connection, thereby the threaded screw 11a of the damper 11 engages with the threaded hole 9c in the bearing seat 8. Similarly, the outer end face of the damper 11 is provided with a threaded opening 11c for engaging with 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. The embodiment shown in the figure is for a bearing assembly / suspension in NDE4, but it should be noted that the damping stiffness of about 106 N / m is substantially similar in DE3 as well. In alternative embodiments, the dampers are cylindrical, but in the preferred embodiment shown in Figure 3, for the rotor 4, there are 20 frustoconical dampers 11 spaced equidistant from each other around the bearing seat 8. Each damper 11 does not contact adjacent dampers 11 in an annular arrangement. The embodiment shown in Figure 3 is for a fourth rotor of a silk reeling device having four rotors. In an alternative embodiment, as shown in Figure 6a, the rotor has approximately 12 to 22 frustoconical dampers 11' that are equidistant from each other.

[0101] Figures 6a and 6b show alternative embodiments of the first bearing assembly 5' and bearing seat 8' for third and fourth rotors, having twelve substantially cylindrical dampers 11 equidistant from one another around the bearing seat 8', each damper 11 having a Shore A hardness of about 55 and a damper ring stiffness of about 106 N / m. The dampers 11 are connected to the bearing seat 8' by screw connections, the threaded screws of the dampers 11 engaging with threaded holes in the bearing seat 8'. Similarly, the other end of the damper 11 has a threaded opening for engaging with bolts for connecting the damper 11 to the rotor housing. The bearing seat 8' further comprises a ball bearing internal steel ring 16' and an aluminum labyrinth ring 10' used to reduce seat mass.

[0102] Referring to Figures 1, 2, and 7, the damper 11 is positioned between the bearing seat 8 and the rotor housing 12 so as to form an annular bearing ring. As shown in Figure 7, the rotor housing 12 of the present invention has a clearance of approximately 14 mm between the bearing assembly 5 and the inner surface of the housing 12 at both NDE4 and DE (not shown). In use, the rotor 1 will be mounted on a high-strength steel wheel at position 13, further along the shaft 2 from the first bearing assembly 5 at NDE4. Referring to Figure 2, the wheels 13a, 13b, 13c, and 13d are positioned at the first end of the shaft 2, and the opposite end is connected to the high-speed motors 30a, 30b, 30c, and 30d by flexible couplings. There is a relatively narrow space between the wheels 13a, 13b, 13c, and 13d. In the example of the fourth rotor shown in Figure 1, the wheel is substantially cylindrical with a weight of approximately 50 kg, a diameter of approximately 332 mm, an outer wall thickness of approximately 25 mm, and a side wall thickness of approximately 15 mm. It should be understood that the outer wall of the wheel has a curved surface, and the side walls are circular and substantially perpendicular to the length of the shaft 2. Referring to Figure 2, typically the wheel diameter of wheel 3 is approximately 314 mm and has a mass of approximately 50 kg. In the preferred embodiment of the fourth rotor, the wheel is approximately 47 mm from the centerline of the NDE bearing assembly 5 and rotates at approximately 6000 RPM to approximately 13000 RPM. The rotor housing 12 supports the bearing and covers the middle portion of the shaft 2. However, for ease of understanding the present invention, the wheel is not shown in Figure 1.

[0103] Referring to Figures 2 and 8a, each rotor housing 12 comprises two symmetrical combination parts, each of which is a semi-cylindrical shell, so that when the two parts are combined, a substantially cylindrical housing 12 is formed. Each half of the rotor housing 12 is held by the adjacent half of the housing by screws 33. Figure 8a also shows screws 11d that hold each damper (not shown) to the rotor housing. The rotor housing 12 further comprises brackets 35 that support tubes and wires. Rubber rings 34 at each end 4, 5 of the rotor housing 12 are used to mount the rotor 1 into the spinner.

[0104] Referring to Figure 8b, in a preferred embodiment, the rotor housing 12 has a wall thickness at the base of the housing 12 that is about 5 mm greater than the upper wall thickness of the rotor housing 12, which effectively lifts the wheel to compensate for the overhang. The internal profile of the rotor housing 12 is asymmetrical. Referring to Figure 8b, the damper 11' at the top of the NDE 4 is extended by about 2.5 mm. The damper 11'' at the bottom of the NDE 4 is 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 action of gravity on the suspension of the bearing 7, the shaft 2, and the wheel 13 within the rotor housing 12, and as a result, the center of gravity of the rotor 1 is in the desired position.

[0105] Referring to Figures 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 desired speed and pre-failure life. These two bearings 7 in DE3 and NDE4 are fitted together in close proximity at each end of the rotor 1 with almost no space between them.

[0106] Referring to Figures 9a, 9b, and 9c, these bearings are ceramic ball bearings 7 between two steel bearing rings 16 and 17. Bearings 7 are selected to withstand loads up to 1400 N. The dynamic load unbalance at the wheel of rotor 4 has been determined to be approximately 560 g·cm, and the configuration of the bearing seat and damper suspension has been optimized to compensate for this unbalance. The angular contact ball bearings 7 have a diameter of approximately 70 mm, although in alternative embodiments, the bearing diameter is one of 60 mm, 65 mm, 70 mm, or 75 mm.

[0107] In the embodiments 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 internal spacer ring 14 between the rotating inner surfaces of ball bearings 7 adjacent to the shaft 2 and an external spacer ring 15 between the rotating inner surfaces of ball bearings 7 furthest from the shaft 2. The ball bearings 7 are held between an internal steel ring 16 mounted on the steel shaft 2 and an external steel ring 17. Each angular contact ball bearing 7 has an angular contact (pressure) angle β that is symmetrical around the center 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] Referring to Figure 9b, when shaft 2 is colder, the shaft will contract to a smaller diameter, so the contact angle β will increase. Referring to Figure 9c, when shaft 2 is warmer, the shaft will stretch to a larger diameter, so the contact angle β will decrease. As shown in Figure 9b, when shaft 2 is cold, it has a smaller diameter Rc and a larger pressure angle β. The outer spacer ring 15 has a width Wc and does not contact either the ball bearing internal steel ring 16 or the ball bearing external steel ring 17, giving it a larger range for ball bearing movement / rattling. However, as shown in Figure 9c, when shaft 2 is warm, it has a larger diameter RH and a smaller pressure angle β. The increased diameter of the shaft also means that the bearing balls 7 push each other toward each other until the outer bearing seat 17 contacts 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] Referring to Figures 9b and 9c, it is found that it is advantageous for the external spacer ring 15 to be shorter than the internal spacer ring 14. Typically, for rotors 3 and 4 with a bearing diameter of 70 mm, the width of the external spacer ring 15 is approximately 61 μm smaller than that of the internal spacer ring 14. For rotor 2 with a bearing diameter of 70 mm, the width of the external spacer ring 15 is approximately 16 μm smaller than that of the internal spacer ring 14.

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

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

[0112] Referring to Figure 10, the rotor 1 further comprises a DE air purge inlet 26 and an NDE air purge inlet 27. The air purge system washes contaminants away from the rotor 1. The rotor 1 further comprises an accelerometer 28 at NDE4 and an accelerometer 29 at DE3, and a pyrometer 30 at DE3 and a pyrometer 31 at NDE4.

[0113] Referring to Figures 10 and 11, the air purge system washes contaminants away from the rotor when the spinner is in use. The rotor also has a water cooling system with a cooling water outlet 24 for removing heated water from the bearing 7. When cleaning the spinner during routine maintenance, it has been found that water can contaminate the bearing 6. Therefore, the present invention further includes a grease labyrinth 38 to prevent water from contaminating the bearing 7 during spinner cleaning, as shown in Figure 11. Figure 11 shows a bearing assembly 6 of DE4 with a rubber ring 34 for mounting the rotor 1 inside the spinner. In the case of DE3 in the figure, there are DE labyrinth grease inlets 38a and NDE labyrinth grease inlets 38b. The NDE labyrinth grease inlet 38b delivers grease to NDE4, which has a similar grease labyrinth. When the rotor is stationary because the spinner is being serviced, the labyrinth 38 is sealed with grease to prevent water from contaminating the bearing 7 during cleaning. By sealing the ring 39, labyrinth grease is prevented from flowing into the bearing 7, which is permanently lubricated by its own grease.

[0114] Referring to Figure 12, a spinning apparatus having four rotors 1a, 1b, 1c, and 1d is used to produce artificial glass fiber (MMVF) when in use. Each rotor 1a, 1b, 1c, and 1d is mounted to rotate around a different substantially horizontal axis, and each rotor 1a, 1b, 1c, and 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 has a suspension ring bearing assembly at the non-driven end having 14 neoprene rubber dampers, and the third and fourth rotors at the non-driven ends have bearing assemblies having 20 neoprene rubber dampers. The second rotor has 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, and 1d has a high-speed motor connected at one end by a flexible coupling, and the wheels are positioned at the opposite ends of rotors 1a, 1b, 1c, and 1d. In the embodiment shown in Figure 12, rotors 1 and 3 rotate counterclockwise, while rotors 2 and 4 rotate clockwise. The wheel of the first rotor 1a has a diameter of approximately 184 mm, the wheel of the second rotor 1b has a diameter of approximately 234 mm, the wheel of the third rotor 1c has a diameter of approximately 314 mm, and the wheel of the fourth rotor 1d has a diameter of approximately 332 mm. The space between the outer surfaces of each mounted wheel varies, with the largest distance being approximately 228 mm between the wheels of the first and fourth rotors 1a and 1d, and the smallest distance being approximately 17 mm between the first and second rotors 1a and 1b.

[0116] Referring to Figure 12, as rotors 1a, 1b, 1c, and 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 MMVF. The melt is then continuously fed into the wheels of the remaining rotors 1b, 1c, and 1d, with fibers being formed and collected each time. Simultaneously, as the melt is passed to the wheels of each successive rotor 1a, 1b, 1c, and 1d, the fibers are removed from the wheels for collection by a high-pressure airflow through the spinner and along the wheels.

[0117] In a typical four-rotor spinning apparatus according to the present invention, wheel 1 produces approximately 5% of the stone wool production per hour, wheel 2 produces approximately 25%, wheel 3 produces approximately 40%, and wheel 4 produces approximately 30%. Manufacturing using the rotors of the present invention can continue for approximately 4000 hours before the ball bearing 7 requires replacement, which is a significant increase over known devices. Durability tests comparing a conventional spinner operating at maximum speed (9300 RPM) with a spinner according to the present invention operating at 13000 RPM, both operating with an unbalance of 560 g·cm, showed an improvement of over 4000 hours from 603 hours. Further tests found a mean interval between failures of approximately 15000 hours for the spinner according to the present invention.

[0118] Referring to Figure 13, simulations were performed for a range of shaft length (Lshaft), shaft diameter (Dshaft), dampering stiffness (Kseat), bearing seat mass (Mseat), and bearing diameter to evaluate the optimal parameters for the improved rotor of the present invention. The effects of these values ​​on bearing life (L10) and rotor shaft displacement (Δx) are shown in the 5D plot.

[0119] Figure 13 is a rear view of a 5D contour plot for rotor 4 at a maximum rotor speed of 13,000 RPM for an NDE seat with a bearing diameter of 70 mm. The height of each peak (L10) represents the bearing life or basic life rating in units of time. The bearing life statistics used are measures of the amount of time in rotation at which 90% of the ball bearing can be expected to remain.

[0120] As shown in Figure 13, the optimal configuration of the present invention enables a maximum bearing life of 3617 hours. The simulation considers a range of shaft diameter (Dshaft) and bearing seat mass (Mseat) plotted together 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 in Figure 13 symbolizes a 3D plot of a function of life (L10) with shaft length (Lshaft) and shaft diameter (Dshaft) as arguments. Each 3D subplot has rubber stiffness (Kseat) and bearing seat mass (Mseat) constants indicated by the position of the 3D subplot.

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

[0122] Referring to Figures 13 and 14, the rubber stiffness (Kseat) at the bearing seat is optimized to be approximately 106 N / m or less. Furthermore, 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 (3 kg), the bearing life is improved. It was also found that when the seat mass (Mseat) is significantly reduced to, for example, approximately 1.5 kg, it becomes necessary to reduce the rubber stiffness (Kseat) at the bearing seat in order to achieve an increased bearing life. It has also been shown that the present invention achieves improved bearing life for a shaft length of 590 mm (Lshaft) and a shaft diameter of 100 mm (Dshaft). Further embodiments of the present invention with a bearing diameter of 60 mm have been investigated, and it has been found that the present invention can achieve a maximum bearing life of 6183 hours.

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

[0124] Within this specification, 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] 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%, and most preferably plus or minus 2%.

[0126] The above embodiments are given only as examples, and those skilled in the art will understand that numerous modifications can be made without departing from the scope of the claims.

Claims

1. A rotor for a silk harvesting device, Rotor housing and First and second bearing assemblies, each bearing assembly comprising at least two ball bearings, each seated within its respective bearing seat, The system comprises a substantially horizontal shaft rotatably mounted between the first bearing assembly and the second bearing assembly, A rotor characterized in that a plurality of elastic dampers are arranged in an annular pattern, and each elastic damper is connected to and engaged with the bearing seat at a first end and connected to and engaged with the inner wall of the rotor housing at a second end.

2. The rotor according to claim 1, wherein each elastic damper is a frustum.

3. The rotor according to claim 1 or 2, comprising a plurality of frustoconical elastic dampers, each damper having a larger diameter on the surface of the damper adjacent to the inner wall of the rotor housing and a smaller diameter on the surface of the damper adjacent to the bearing seat, and / or the damper or each damper is a rubber damper, a silicone damper, or a neoprene rubber damper.

4. The rotor according to any one of claims 1 to 3, wherein the plurality of elastic dampers are adapted to the operating rotational speed of the rotor between 4,000 RPM and 13,000 RPM.

5. The rotor according to any one of claims 1 to 4, 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.

6. The rotor according to any one of claims 1 to 5, wherein the damper or each damper has a Shore A hardness of 55.

7. The rotor according to any one of claims 1 to 6, wherein each damper is provided with a threaded screw for releasably connecting to the bearing seat, and / or each damper is provided with a threaded opening for releasably connecting to the screw through the rotor housing.

8. A rotor according to any one of claims 1 to 7, comprising approximately 10 to approximately 24 frustoconical dampers arranged in a ring.

9. The rotor according to any one of claims 1 to 8, wherein the bearing or each bearing is a hybrid angular contact ball bearing, and / or the inner diameter of the ball bearing or each ball bearing is 60 mm to 75 mm, and / or the bearing is made of ceramic material, and / or the distance between the two angular contact bearings is 20 mm.

10. The rotor according to any one of claims 1 to 9, wherein the clearance between the annular bearing seat and the inner surface of the rotor housing is 14 mm, and / or the two angular contact ball bearings are separated by an internal axial spacer ring and an external axial spacer ring.

11. The rotor according to claim 10, wherein the width of the external spacer ring is smaller than the width of the internal spacer ring.

12. The rotor according to claim 10 or 11, wherein the width of the external spacer ring is 16 μm to 61 μm smaller than the width of the internal spacer ring.

13. The rotor according to any one of claims 1 to 12, wherein the relationship between the shaft diameter (Dshaft) and the shaft length (Lshaft) is defined as Dshaft(Lshaft) ≥ 0.12 * Lshaft - 32 mm for shaft lengths in the range of 101 mm to 1325 mm and shaft diameters of 20 mm or more.

14. The rotor according to any one of claims 1 to 13, wherein the external cross-sectional diameter of the shaft is 100 mm, and / or the shaft has a bearing seat diameter of 70 mm, and / or the length of the shaft is 955 mm, and / or the length of the shaft between the center point of the first bearing assembly and the center point of the second bearing assembly is 590 mm.

15. The rotor according to any one of claims 1 to 14, wherein the weight of the bearing seat or each bearing seat is 3 kg or less.

16. The rotor according to any one of claims 1 to 15, wherein the dampers are equidistant from each other substantially around an annular bearing assembly.

17. A yarn picking device comprising a set of at least three rotors according to any one of claims 1 to 16, wherein each rotor is mounted to rotate around a different substantially horizontal axis, and is arranged such that, as the rotors are rotating, melt poured onto the circumferential surface of a first rotor in the set is continuously fed onto the circumferential surfaces of each subsequent rotor, and fibers are discharged from the rotors.

18. A method for producing artificial glass fiber (MMVF), A step of providing a silk-collecting device comprising a set of at least three rotors according to any one of claims 1 to 16, each rotor mounted to rotate around a different substantially horizontal axis, wherein each rotor has a driving means; The steps of rotating the rotor, A step of providing a mineral melt for forming an artificial glass fiber (MMVF), wherein the melt is poured onto the periphery of a first rotor, A method comprising the step of collecting the formed fibers.

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