Diagnostics practice of heavy duty high speed gear transmissions/Dideles galios krumpliniu pavaru diagnostikos praktika.

Barzdaitis, V. ; Mazeika, P.

1. Introduction

The gear transmission is a high frequency mechanical vibration system with the gear mesh as the main vibration source. The low rotational speed heavy duty gear power transmission have been modeled and tested in situ to protect unexpected failures [1, 2, 5]. The high rotational speed heavy duty gearing diagnostics is important in modern energy generating machines. In case of complex rotary machines, it is very complicated to trace all sources of vibrations (e.g. internal defects) [6]. The rotor systems with adaptive and sleeve sliding-friction bearings is analyzed in the paper [7]. Due to vibrations of gears, the teeth meshing inaccuracies generate dynamic forces and early wear teeth. So far, the effects on gears loading have been well studied, but an issue that the intensity of vibration acceleration level and teeth meshing shocks--how severe is the problem--has scarcely been studied for practical purposes. The gears vibration acceleration parameter was used to diagnostics and evaluation of teeth meshing problem. This article is dedicated for condition monitoring of the gear power transmissions, protection and failures diagnostics through vibration parameters measured in situ.

2. The objects of research

The modern, high efficiency power generating machines with gas and steam turbines have high rotational speed of rotors (8000 rpm and more). Rotational speeds of the electric generators rotors are 1500 rpm--3600 rpm. The specificity of these machines design included various types of gear transmissions that reduce the rotational speeds of turbines to electric generators.

[FIGURE 1 OMITTED]

The turbo units with power ratings from 700 kW to 10000 kW with gear drivers are experimentally tested in situ using bearing housings absolute vibration parameters measured with seismic transducers--pjezoaccelerometers. The gear power transmissions usually consist of two or more gears meshed for the purpose of transmitting motion (power) from turbine rotor to electric generator rotor. The motion parameters are: P--power, T--rotor torque, n--rotational speed of the rotor or [omega]-rotor angular velocity.

3. Vibration of the helical gear reducer

Kinematical scheme of turbo unit with steam turbine, ordinary double-helical gear train and electric generator set is shown in Fig. 1.

The two different power generating turbounits have analogous kinematic schemes. The 700 kW power steam turbine (rotor running on four-lobe journal bearings 1 and 2) rotational speed is [n.sub.ST1] = 8000 rpm (133.33 Hz) and the 1250 kW power steam turbine rotational speed is [n.sub.ST2] = 10500 rpm (175 Hz). Rotational speed of the both electric generators rotors (running on antifriction bearings 7 and 8) is low [n.sub.EG] = 1500 rpm (25 Hz). The steam turbine rotor is connected by a coupling 12 to the ordinary double-helical gear train with the speed ratio 1:5.3448 (pinion [z.sub.1] = 29, gear [z.sub.2] = 155) and the second unit gear transmission speed ratio is 1:6.857 (pinion [z.sub.1] = 21, gear [z.sub.2] = 144) that reduces turbine rotor high rotational speed to 1500 rpm of the generator rotor 11. The motion parameters of the gear train driven rotors: of the 700 kW power machine - [T.sub.1] = 4460 Nm, [[omega].sub.1] = 157 rad/s and of the 1250 kW power machine - [T.sub.2] = 7960 Nm, [[omega].sub.2] = 157 rad/s.

Vibration acceleration spectrum of the 700 kW power gear transmission bearing housing is shown in Fig. 2. The gears [z.sub.1] = 29, [z.sub.2] = 155 mesh generate high frequency (29-133.33 Hz = 3880 Hz) shocks with severe vibration acceleration amplitudes (143 m/[s.sup.2]) of 4th bearing, Fig. 2. The measured vibration intensity is too severe for long term safe operation mode of the turbounit. The increased radial gaps in journal bearings 3-6 or damages of white metal in bearings (rubbing) generated transmission error in gear mesh.

Frequency of the 1250 kW power turbounit gear mesh is (21-175 Hz = 3600 Hz) with high level vibration acceleration amplitude of 4th bearing measured in vertical direction. The plots of root mean square vibration acceleration value [a.sub.rms] = 113.8 m/[s.sup.2] (161 dB) and maximum peak value [a.sub.max_peak] = 227.3 m/[s.sup.2] (167 dB) are shown in Fig. 3. The high vibration intensity is too high for safe continuous operation of the turbounit. The damaged journal bearing generates transmission error in gear mesh and generates high vibration amplitude of gear mesh frequency and sidebands.

[FIGURE 2 OMITTED]

[FIGURE 3 OMITTED]

4. Vibration of the high rotational speed epicyclical gear train

Planetary gear trains, also referred as epicyclic gear trains, are those in which one or more gears orbit about the central axis of the train. Thus, they differ from an ordinary parallel shaft gear trains by having a moving axis or axes: sun pinion gear, planet gear, planet carrier, internal toothed ring gear--annulus, Fig. 4. The most significant advantage of epicyclic gears is that the input torque is distributed to all of the planets gears with optimum utilization of space and minimum weight due to effective load distribution. The torque is balancing equally between the planets gears. The radial forces of each gear teeth mesh offset each other that allow achieve compact design with the best load equalization at tooth contact points. The double helical epicyclic gears are especially suitable for high powers and speeds (power range 1 37 MW, high speed shafts 3000-19000 rpm, low speed shaft 1500 or 1800 rpm), e.g. gas or steam turbine driven el. generators. They eliminate axial thrust, prevent tilting moments and ensure silent running owing to the larger contact ratio of meshing teeth, improved distribution of heat in the tooth area. Usually all shafts run in plain bearings or multilobe bearings.

Epicyclic gear units can operate at up to 1% higher efficiency than a parallel shaft gear unit, and are inherently quieter than parallel shaft gear units, with lower noise and vibration [6].

The 9.5 MW power steam turbine-electric generator unit is additional power module of the 25 MW power gas turbo unit, Fig. 5. The epicyclic gear train is integral module of this power generating machine. The sun gear [z.sub.1] = 29 is connected to the steam turbine high speed rotor with rotational speed 11224 rpm (187 Hz). The planet carrier is stationary connected to the gear train case. The annulus has [z.sub.2] = 217 teeth and connected to the four poles electric generator rotor with 1500 rpm. The motion parameters of the 9500 kW power gear train driven rotor are: T = 60510 Nm, [omega] = 157 rad/s, gear ratio u = 7.4813.

[FIGURE 4 OMITTED]

[FIGURE 5 OMITTED]

The gear case absolute vibrations were measured in horizontal-radial direction with piezo accelerometers stationary attached at planet carrier fixed location on the gear case point, as shown in Fig. 6.

Root mean square of the steam turbine 2nd bearing housing horizontal vibration velocity value is low 0.5 mm/s, Fig. 5. The vibration velocity spectrum has scarce information about condition of the gearing because the vibration intensity is too low measured up to 1000 Hz interval: el. generator rotor driven by annulus vibration amplitude 0.29 mm/s at synchronous frequency 25 Hz, the sun gear rotation frequency 187 Hz vibration velocity amplitude is low 0.34 mm/s and dominated in the spectrum.

Root mean square of the epicyclic gear train case absolute horizontal vibration velocity value is 0.8 mm/s and has scarce information about vibration of gears, Fig. 5. The vibration velocity comprises the el. generator rotational frequency 25 Hz amplitude 0.38 mm/s, the planet gear rotational frequency is 57.5 Hz amplitude 0.61 mm/s, 373.7 Hz amplitude 0.1 mm/s.

Acceleration spectrum of the epicyclical gear train case horizontal vibration is more informative in diagnostics and shown in Fig. 7. The gear mesh frequency is high ~5420 Hz with acceleration amplitude 52 m/[s.sup.2] and dominated in the spectrum (29-187 = 5423 Hz). This difference indicated some inaccuracies of stationary monitoring system, as shown in Fig. 5 (2.1 g vibration acceleration).

[FIGURE 6 OMITTED]

[FIGURE 7 OMITTED]

[FIGURE 8 OMITTED]

Vibration acceleration amplitude of the gear mesh frequency 5420 Hz, measured in vertical direction, is few times less in comparison with horizontal direction, because dynamic stiffness of the gear case in horizontal direction is less in comparison with the vertical direction.

As diagnostics practice showed that not only the antifriction bearings technical condition can be evaluated according plot in Fig. 8, but partially the gearings teeth meshing conditions [7]. The ordinary double-helical gear generates high teeth meshing high shocks in comparison with epicyclical gear and can be evaluated partially using data in Fig. 8.

5. Conclusion

The ordinary helical gear train has few times higher vibration acceleration level in comparison with the epicyclical gear train.

The peculiarity is that each gear box must be evaluated separately using condition monitoring system with vibration acceleration and frequency parameters.

The vibration measurement location is decisive for evaluation of gear mesh technical condition and could be chosen as the most sensitive point on the gearbox for location of the transducers. The measured vibration values as gear mesh frequency [f.sub.mf] vibration acceleration amplitudes a, root mean square values [a.sub.rms] and maximum vibration acceleration values [a.sub.peak] are the most informative in diagnostics and can be evaluated similar as high frequency vibration intensity of antifriction bearings.

Received December 01, 2009 Accepted February 05, 2010

References

[1.] Barzdaitis, V., Bogdevi?ius, M., Didziokas, R., Mazeika, P. Modeling and testing of gear power transmission. -Proceedings, 4th International Symposium on Stability Control of Rotating Machinery. Calgary, Alberta Canada 27-31 August, 2007, Minden Nevada, USA, p.321-330.

[2.] Mazeika, P. Diagnostics and Failures Prevention Researches of Rotors with Rolling Element Bearings. Summary of Doctoral Dissertation. -Kaunas: Technologija, 2008.-36p.

[3.] Vasylius, M., Didziokas, R., Mazeika, P., Barzdaitis, V. The rotating system vibration and diagnostics. -Mechanika. -Kaunas: Technologija, 2008, Nr.4(72), p.54-58.

[4.] Juzenas, E., Jonusas, R., Juzenas, K. Research of complex rotary systems vibrocondition based on analysis of dynamical processes and spectrum of vibrations. -Mechanika.-Kaunas: Technologija, 2008, Nr.1(69), p.42-45.

[5.] Vekteris, V., Cereska, A., Jurevicius, M., Striska, V. Experimental research of rotor axis revolution orbit in rotor systems with adaptive and sleeve sliding-friction bearings.-Mechanika.-Kaunas: Technologija, 2008, Nr.2(70), p.38-42.

[6.] Epicyclic Gear train; Voith Turbo BHS Getriebe GmbH, 2007-10-02, www.bhs-getriebe.de, info@bhsgetriebe.de).

[7.] Table of Criteria Bearing measurements (10-10000 Hz), from Canadian Government Specification CDA/MSNVSH 107: Vibration Limits for Maintanace. Bearing Severity Chart. Bruel & Kjear, QH 0031, 2000.

V. Barzdaitis *, P. Mazeika **

* Kaunas University of Technology, A. Mickeviciaus str. 37, 44244 Kaunas, Lithuania, E-mail: vytautas.barzdaitis@ktu.lt

** Klaipeda University Mechatronic Science Institute, Bijunu_str. 17, 91225 Klaipeda, Lithuania, E-mail: pranasmazeika@centras.lt