PNEUMATIC HIGH SPEED SPINDLE WITH AIR BEARINGS
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1 PNEUMATIC HIGH SPEED SPINDLE WITH AIR BEARINGS Terenziano RAPARELLI, Federico COLOMBO and Rodrigo VILLAVICENCIO Department of Mechanics, Politecnico di Torino Corso Duca degli Abruzzi 24, Torino, Italy ( ABSTRACT This article describes different experimental tests performed on a pneumatic spindle with gas bearings in order to evaluate its maximum rotational speed and stability. In particular the rotor unbalance response, the power losses calculated from deceleration tests, the thermal transient due to viscous losses are calculated and analyzed. KEY WORDS Pneumatic Spindle, Air Bearings, Feed hole, Thermal transient INTRODUCTION These applications are related both to high precision devices, e.g. for measuring machines, both to high speed rotating machines, e.g. high speed spindles for operations of finishing or drilling. During the last fifty years the use of no contact air bearings in rotating machines has been the object of interest by part of the researches in order to obtain bearings that support rotational speeds out of the range of conventional ball bearings [1]. Limitations on the maximum rotational speed are represented by centrifugal forces and by the well-know whirl instability of gas bearings. Nevertheless gas bearings enable to reach higher rotational speeds than rolling bearings, compared at the same diameter. Focusing the attention on the high rotational speeds, a broad spectrum of applications demand the use of gas bearings, in manufacturing industry (high speed machining, printed circuit boards drilling, micro-milling and wafer dicing) and not only (high speed compressors and turbines). In the first case the high speed is aimed to obtain high quality surface finishing or the correct tangential cutting speed with micro-tools; in the second case it is due to a reduction of the dimensions of the turbomachines. The drawbacks of pneumatic bearings can be summarized in the following phenomena: the air-hammer [4], the unstable whirl [5][6], the relative low damping. All these problems must be solved to design such systems and obtain a stable operation of the rotor to contain its dynamic runout.
2 PROTOTYPE AND TEST BENCH Figure 2 depicts the cross section of rotor (2), which is sustained by two radial bearings (4) of axial length 100 mm and a double thrust bearing (3) composed by three rings. The bearings are supplied by means of channels drilled in the housing (1) and in the three rings of the thrust bearing. The thrust bearing has inner and outer radius 49 mm and 79 mm respectively. The thickness of the central ring defines the axial air gap. The nominal radial and axial clearance are 27 µm and 19 µm respectively. The rotor is accelerated by a pneumatic turbine (5) machined on the rotor itself. Figure 1: Sketch of the pneumatic spindle. In this paper, a completely pneumatic spindle (Figure 1) designed and realized at the Mechanical Department of Politecnico di Torino [2] was tested experimentally. The stability of the spindle was verified monitoring the dynamic runout by means of displacement sensors facing the rotor and positioned along perpendicular radial directions. The bearings air consumption [3] and the external surface temperature were registered at different rotating speeds. The air turbine used to accelerate the rotor was also characterized. Particular attention was given to the thermal analysis to evaluate the effects of temperature on the air clearance and prevent any rotor grip. Figure 2: Prototype pneumatic spindle. Figure 3: Transducers and Data Acquisition system. The test bench designed to measure the performance of the pneumatic spindle was endowed of a data acquisition system and of pneumatic supply lines. Displacement capacitive transducers, see Figure 3 are arranged radially and axially facing the rotor. Four radial transducers are placed in two planes (close to each bearing) along X and Y directions. In Figure 3 are represented the measuring planes on which sensors (1) and (2) are set on the turbine side and sensors (3) and (4) on the thrust side. Sensor (5) is set along the axial direction on the thrust side. The sensors are connected to the amplifiers (6) and to the signal conditioner (7), whose analog output is detected by means of the data acquisition system. An optical tachometer (8) is placed in front of the spindle along the axial direction. The pulse train signal of the sensor is sent to counters that calculate the spindle rotational frequency. A series of
3 thermocouples type T (9) are disposed on the carter surface to measure its external temperature distribution; their signal is amplified by means of amplifier (10). The temperature of the discharged air is also measured next to the axial and radial bearings. The pneumatic system (Figure 10) consists of two supply lines: one for the turbine and the other for the bearings. This last is endowed of a 100 l tank with non-return valves, which is inserted downstream the pressure regulator. It can be useful in case the supply line fails to prevent a spindle damage. The output signal of the displacement transducers is acquired with a 50 khz sampling rate. The signals of two perpendicular sensors are composed in an XY plot to visualize the spindle orbits on the measuring planes. In this way the conical and cylindrical modes are visualized. By means of a FFT algorithm the Fourier spectrum of the signal is obtained in order to study the unbalance spindle response and verify the absence of sub-synchronous whirl. Six thermocouples were set in correspondence of the following points (see the Figure 4): 1-turbine housing, 2-air discharge next to the turbine, 3-measuring plane on turbine side, 4-measuring plane on thrust side 5-rotor flange, 6-air discharge next to the thrust bearing. 1) the unbalance spindle response, 2) the bearings air consumption, 3) deceleration tests, 4) temperature tests, 5) the turbine characterization. The measuring of the rotor and bearings dimensions was intended to evaluate the radial and axial air clearance, to which the bearing stiffness, the viscous losses and the spindle stability during operation are very sensitive. The mean values obtained after many dimensional tests are reported in Table 1: Table 1: Dimension of bearings and air gaps. Rotor dia. on turbine side mm Rotor dia. on thrust side mm Bushing dia. on turbine side mm Bushing dia. on thrust side mm Radial gap on turbine side 27 µm Radial gap on thrust side 27 µm Thickness of spindle flange mm Thickness of central ring mm Axial gap 19 µm Figure 4: Thermocouple measuring points. EXPERIMENTAL ACTIVITY The experimental activity involved beforehand the measure of the air gaps, then were carried out: Figure 5: Deceleration test. To calculate the friction torque, deceleration tests were carried out. The deceleration was started at rpm by closing the spheric valve that intercepts the turbine supply line. Figure 5 displays the spindle rotational speed vs time, while Figure 6 plots the torque due to
4 viscous losses on bearings, that was obtained as a function of the rotational speed considering the mass moment of inertia of the spindle. The spindle orbits were measured every 1000 rpm starting from 0 to rpm. The spindle during operation suffers from centrifugal expansion; for this reason its surface approaches to the sensors, that are fixed to the housing. This displacement is visible in the spindle orbits show in, Figure 8 and Figure 9. Figure 6: Friction viscous torque. The temperature trends vs time are reported in Figure 7. They refer to a gauge supply pressure of 5 bar to the bearings and an initial rotational speed of rpm. After 110 minutes the spindle was shut down. It was observed that the temperature increases moving from the turbine to the thrust bearing, where the tangential speeds are higher. After the spindle shut down the temperature becomes uniform and begins to decrease, with the exception of the turbine housing, that presents an initial increase of temperature. Figure 8: Spindle orbits measured on the turbine side plane at different rotational speeds; (a)non filtered, (b) filtered. Figure 7: Curves of temperature vs. time. The driving turbine was characterized as a function of the rotational speed. Its air consumption was measured with a rotameter and the pressure was detected at the input and output of the turbine, Figure 10. In Table 2 it is shown the correspondence between the air consumption and the supply pressure of the turbine and the steady rotational speed without external payload. In these conditions it is calculated the power absorbed by the turbine, that is equal to the power dissipated.
5 Figure 10: Pneumatic system used to characterize the driving turbine. Figure 9: Spindle orbits measured on the thrust side plane at different rotational speeds; (a)non filtered, (b) filtered. in Figure 8 for the measuring plane on the turbine side and in Figure 9 for the measuring plane on the thrust side. The figures compare the original signal with the signal obtained with a low pass filter. The dynamic runout is confined under 7 µm. In Figure 11 the amplitudes of the runout measured by the four radial sensors are plotted against the rotational speed. The first critical speed is about rpm. The effect of centrifugal forces is visible in Figure 12, where the mean displacement values detected from radial sensors is plotted against the rotor speed. In Figure 12 it is represented the displacement the rotor center around the mean value. Table 2: Power Consumption rpm turbine air gauge press power flow(l/min)anr (bar) (kw) RESULTS The orbits of the spindle during rotation are plotted Figure 11: Amplitude displacement.
6 Figure 14: Mode of rotation at (rpm). Figure 12: Displacement of Y axis around the center of the orbit (turbine side). The unbalance response evidences both cylindrical and conical modes. In Figure 13 and Figure 14 are represented examples of conical modes at rpm and rpm respectively. Figure 15: Waterfall Diagram. The temperature of the rotor increase during operation due to viscous losses in the air gap. It remains in a safety range in which the spindle can operate for a long time without problems. CONCLUSIONS Figure 13: Mode of rotation at (rpm). The waterfall diagram of Figure 15 shows that the unbalance response is only synchronous and the subsynchronous whirl is absent in the tested range of speed. An experimental activity was developed to evaluate the performance of a turbo pneumatic spindle with gas bearings. In synthesis the following conclusions can be listed:
7 the spindle rotates in stable conditions up to rpm; the spindle temperature in steady conditions at rpm remains under 75 o C and the thermal effects due to viscous actions do not compromise the operation of the spindle; the friction torque of gas bearings is very low (around 0.13 Nm at the maximum speed) due to the low air viscosity. REFERENCES [1] Krzysztof, Czolczynski, Rotordynamics of gaslubricated journal bearing systems, New York. (1999), Springer. [2] Belforte G., Raparelli T., Viktorov V., Trivella A., Colombo F., An experimental study of highspeed rotor supported by air bearings: test rig and first experimental results, Tribology International, 2006, 39-8, pp [3] Belforte G., Raparelli T., Viktorov V., Trivella A., Discharge coefficients of orifice type restrictor for aerostatic bearings, Tribology International, 2007, 40, pp [4] Tauler H.M.,Stowell T.B., Pneumatic hammer in an externally pressurized orifice-compensated air journal bearing, Tribology International, 2003, 36, pp [5] Waumans T., Peirs J., Reynaerts D., Al-Bender F., On the dynamic stability of high-speed gas bearings: Stability study and experimental validation, 2011, Laboratory Soete [6] Brzeski L.,Kazimierski Z., Experimental investigation of precision spindles equipped with high stiffness gas journal bearings, Precision Engineering, 2003, 23, pp
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