PRÜFTECHNIK for turbomachinery

No. 13 Focus: Turbomachinery the service magazine of the PRÜFTECHNIK Group PRÜFTECHNIK News PRÜFTECHNIK for turbomachinery The new version of OMNITREND now lets you evaluate and fully analyze shaft vibrations in power plant turbogenerator sets, industrial turbo-generator sets, gas turbine sets, compressor systems and other machinery in a standards-compliant manner. Mobile measurement and diagnosis of shaft vibrations can be readily accomplished with the 2-channel VIBXPERT data collector and signal analyzer, which is connected Condition Monitoring Service Shaft vibrations in a turbo-generator Dr. Edwin Becker to protection monitoring systems or to temporarily mounted displacement sensors. This issue is all about turbomachinery and how new Condition Monitoring approaches make your conditionbased maintenance program even more effective through the analysis of shaft vibrations. We hope you find this issue interesting and enjoyable and will be glad to answer any questions you may have. In this issue: PRÜFTECHNIK for turbomachinery Strong shaft vibrations in a turbo-generator set Evaluating shaft vibrations in turbomachinery Determining turbine alignment targets Alignment of a steam turbine Measuring and evaluating shaft movement: part 1 part 8 News 1 Power plant turbo-generator sets should run 24/7 not only to keep power flowing but also to improve the machinery s life expectancy, which is shortened every time machines are started up and shut down. While it is unfortunate when a turbo-generator set needs to be shut down for maintenance or due to excessive vibration, it is entirely futile if no faults are subsequently found. PRÜFTECHNIK was contracted to conduct a vibration analysis on a 55-MW turbo-generator set. This job came about due to a steady rise in shaft vibrations at the Drive-end (DE) bearing of the generator. The cause was suspected to be vibrations from the turbo gear unit since the gearbox contained corrected tooth edge damage. Would a new gear set need to be ordered? Dr. Becker handled this assignment himself since he is familiar with turbo gear units from earlier jobs. This was also a good opportunity to test the new VIBXPERT and OMNI- TREND functions. The different machine analyses were combined. On the one hand, measurements were taken directly at typical housing measuring points using piezoelectric accelerometers. In addition, shaft vibrations were measured using a Bently shaft vibration protection system installed in the unit. Unfortunately, there were no vibration sensors mounted on the gearbox. The gear condition was evaluated using accelerometers. The first set of analyses showed that the gear vibrations were by no means unusual and that no gear mesh faults occurred under load change. On the other hand, directional forces between the generator and gearbox were found, as well as elliptical orbits on both generator sides. In fact, on the DE side, the amplitudes were impermissibly high at 123 µm. Had the machinery alignment changed? This was checked in a series of specific examinations several days later during a scheduled shutdown. The visual inspection showed unusual contact patterns in the double engagement gear coupling and slight tarnishing on the generator Non-drive-end (NDE) bearing. The alignment had changed only slightly, and the small vertical deviation that was found was corrected. Was it possible that the new gear coupling installed during the last shutdown was the cause? Fig. 1: Unusual contact patterns in the double engagement gear coupling.

PRÜFTECHNIK investigated this when the turbo-generator set was recommissioned. They performed additional comparative measurements, which could also be used to check run-out. Several sensors, measurement results and the build-up of pressure in hydrodynamic bearings are shown in Figs. 2 6. The shaft vibrations measured at the generator NDE bearing were conspicuous. The rotating shaft was virtually glued to the upper bearing shell. The associated time waveforms in Fig. 4 show that the harmonic motion of the rotor was disturbed. Time waveforms for the generator DE bearing are shown in Fig. 6. They provide information on time-based shaft vibrations before and after correction of the alignment. By chance, the turbo-generator set had to be shut down again due to a controller problem. I quickly activated the recording function in VIBXPERT, reported Dr. Becker. During ramp down, it became apparent that the generator was operating at a overcritical level. This explained why the elliptical orbits were so pronounced on both generator sides. This type of rotor flexibility needs to be taken into account in the alignment targets. Only why did the system run better before? Does the current gear coupling lack the necessary crowning to balance out the forces between the coupled machines? Indeed the coupling did not have crowned outer toothing and was thus unable to correctly align the displacements and eccentricities of the coupled machines. A crowned coupling of the same design was ordered and installed. Measurements taken six months later showed that the new crowned tooth coupling eliminated the disturbing vibrations and the drive ran smoothly. Fig. 2: Accelerometer on gearbox. Fig. 3: Measurement signals taken from the Bently rack and recorded with VIBXPERT. Fig. 4: The time waveform of shaft vibration on the generator NDE bearing in +45 and -45 directions, with the associated frequency spectra. Fig. 5: Pressure buildup in a hydrodynamic journal bearing. Fig. 7: Alignment that takes rotor flexibility into account. Fig. 6: Time waveforms of shaft vibration from the generator DE bearing, before and after alignment. 2

Condition Monitoring Background Information Evaluating shaft vibrations in turbomachinery Misel Tanasijevic Turbomachines are highly complex systems used in power plants. They have a relatively high machine value, exhibit a certain risk potential and their failure can result in high secondary costs. For this reason, turbo-generator sets are often equipped with a vibration protection monitoring system to obtain timely information on when to shut down the machine. When the threshold parameters are set up correctly, these protection systems detect slow changes in properties such as imbalance, in dynamic marginal conditions (bearings, screw connections) and in operating conditions/ processes. Diagnostic vibration monitoring determines dependencies on operating parameters, transient operating states, orbits, frequencies and order analyses, and uses correlation methods to identify changes early on. Standardized guidelines and/or norms from the European and American sector serve as decision-making criteria. Their main difference lies in the evaluation variables applied. While API standards use peak-to-peak values as parameters for amplitude evaluation, the European norms like to employ s max values. The applicable standards and guidelines for evaluating shaft vibrations are shown in Fig. 1. The standards assess conditions by assigning them to one of four zones: Zone A: good Zone B: allowable Zone C: just permissible Zone D: not permissible Machines with rotating masses Shaft vibrations These evaluations only apply to continuous operation and not to transient conditions. Condition Monitoring differentiates between machine protection, vibration monitoring and vibration diagnosis. While the protection and monitoring methods are well described in the standards listed here, comprehensive Vibrations rpm-range D additional damage occurs C restricted operation B continuous operation without restrictions A recently put into operation rpm Fig. 2: Example of general evaluation criteria. specifications for diagnosis methods do not yet exist. Initial steps have been taken in VDI 3839, Sheet 3, but these have not yet been finalized. Figures 3 5 show several nomograms from the DIN-ISO standards. If you know the rpm of the machine, it is relatively easy to determine how high the shaft vibration amplitudes are permitted to be. The standards, which are available from the German Beuth-Verlag, also provide information and make recommendations on setting warning and switchoff criteria. S p-p relative [µm] rpm [min -1 ] Fig. 3: Recommended limit values for relative shaft vibrations (stationary steam turbines and generators above 50 MW). Relative shaft vibration displacement, peak-to-peak in µm S p-p relative [µm] rpm x1000 [min -1 ] Fig. 4: Recommended alarm thresholds for shaft vibration in coupled industrial machines. S p-p relative [µm] DIN/ISO 7919-2 DIN/ISO 7919-3 DIN/ISO 7919-4 Manufacturer acceptance Measurement during acceptance Operation monitoring API 541 API 546 API 611 API 612 API 617 VDI 2059 DIN/ISO 7919-1 DIN/ISO 7919-2 DIN/ISO 7919-3 DIN/ISO 7919-4 VDI 2059 DIN/ISO 7919-1 DIN/ISO 7919-2 DIN/ISO 7919-3 DIN/ISO 7919-4 rpm x1000 [min -1 ] 3 Fig. 1: Some applicable standards for shaft vibration in turbomachinery. Fig. 5: Recommended alarm thresholds for shaft vibration in gas turbine sets.

Alignment Application Determining the alignment targets of a turbine Dirk Günther To be able to correctly align a turbine system, its thermal growth characteristics must be known and taken into consideration. This is the only way to ensure low-vibration operation and a long component service life. In critical turbine systems, online continuous monitoring of the alignment condition may be wise. This measurement detects foundation and system movements to enable appropriate intervention early on. PERMALIGN measures alignment changes PERMALIGN was developed by PRÜFTECHNIK specifically for these applications and can be used with compressors, steam and gas turbines, and water cooling systems. There is also a special model for the chemical, oil and gas industry with explosion protection (Ex ib IIc T4 Zone 1). Four-machine train with increased vibrations In a four-machine train consisting of a steam turbine, two compressors (HP, LP) and an expander (Fig. 2), repeat occurrence of bearing damage and raised vibration levels led to the conclusion that misalignment existed due to a failure to take thermal growth into account. This was a task for PERMALIGN and the PRÜFTECHNIK Machinery Service Team. With PERMALIGN, relative machine displacements can be recorded during operation in four degrees of freedom using a roof prism. Axial displacements and thermal growth can be measured Fig. 1: PERMALIGN roof prism with mounting bracket. directly with triple prisms at a distance of up to 10 meters with micrometer resolution. The service call PERMALIGN was installed using stable brackets with the system running (Fig. 3). Then the machinery was ramped down and switched off. The measurement data recorded over this period was ideal for determining the optimal alignment targets, according to which the machine train was then Fig. 3: PERMALIGN systems mounted on the machine train for measuring positional changes in the vertical and horizontal directions. aligned. During the subsequent machinery startup, PERMALIGN sensors remained installed to compare the displacements with the expected values. The result By using alignment targets that take thermal growth into account, the vibration values at operating temperature were considerably reduced. The operator now has an optimally aligned machine train, which will have a positive effect on component service life. 4 Fig. 2: Measurement of a four-machine train. Preview The next issue will focus on generator systems: Wear monitoring in combined heat and power plants Acceptance measurements on a ship generator Excessively high vibrations in a gas engine Applicable vibration standards for generators

5 Alignment Application Alignment of a steam turbine Bernardo Quintana Steam turbines are important machines for power generation. More than half of electricity produced worldwide is generated using steam turbines. They exist in a wide range of sizes. Typical power outputs for small industrial applications are 2 3 MW, while large coalburning or nuclear power plants have turbines that generate up to 1 GW. Steam turbine efficiency is closely related to the air gap between the turbine components and the rotor. This is why the installation position of the rotor relative to the internal housing is extremely important. It should be checked closely, not only when first installed but also when the machine is overhauled. The CENTRALIGN Ultra measurement system developed by PRÜFTECH- NIK has been in use to align turbine shells to rotors for many years around the world. Faster than piano wire or dummy shafts Compared to conventional methods like piano wire and dummy shafts, this method saves a lot of time and delivers unequaled accuracy with its laser-optical measurement procedure. With CENTRALIGN Ultra, internal turbine elements like bearing shells, turbine casings, guide vanes and internal shells can be aligned with a high degree of precision. The use of a laser extends the working range to 40 m without sag, and the sensor measures with a resolution of 1 µm. An additional control sensor permanently monitors the measurement setup for changes and compensates laser drift. The service call PRÜFTECHNIK was contracted to align half-shells to a rotor during routine maintenance of a turbine. After the casing was opened, the rotor was removed and transported to a partner who repaired or replaced the individual blades and then balanced the entire rotor. In parallel, the turbine elements and labyrinth seals were renewed on the Laser Measuring sensor with precision holder Fig. 1: CENTRALIGN Ultra measurement system setup. turbine system. Finally, the CENTRA- LIGN Ultra system and Machinery Service got to work. The laser and the control sensor were mounted and adjusted outside of the turbine bearings as shown in Fig. 1. The laser beam acts like a reference line for measuring each individual turbine element just like a piano wire would. ROTALIGN Ultra computer with wireless signal transmission Control sensor Each half-shell was measured with the patented sensor measurement equipment from PRÜFTECHNIK and the data was transmitted remotely to the ROTA- LIGN Ultra computer. After all measuring locations were recorded, two bearings or oil rings were selected in ROTALIGN Ultra as a reference and all horizontal and vertical deviations rela- Fig. 2: The measurement values of the individual half-shells are remotely transmitted to the computer.

tive to this reference were measured. The correction values for the bearings of the half-shells were computed from these deviations, enabling the installers to implement these promptly. Fig. 3: The precision laser mounted outside of the turbine bearings. Fig. 5: The aligned steam turbine with the control sensor in the foreground. Fig. 4: Mounting of a new baffle. 6 CENTRALIGN Ultra advantages Precise laser equipment with a resolution of 1 µm No lengthy installations and adjustments of piano wire and dummy shafts The laser beam is not subject to the force of gravity and does not sag Individual segments can be worked on immediately after the measurement data is taken since they are fully accessible Measurement data are immediately available in electronic form and target values for rotor sag can be taken into account using the ALIGNMENT CENTER software Fig. 6: Vertical (top) and horizontal (bottom) measurement results as displayed by the ALIGN- MENT CENTER PC software. Fig. 7: Finally, the overhauled rotor can be installed back in the turbine casing.

7 Condition Monitoring Basics Measuring and evaluating shaft movement Non-contact displacement sensors (part 1) Dr. Edwin Becker Shaft movements take the form of (slow) displacements and/or (rapid) vibrations of the turning rotor shafts in axial and radial measurement directions. Vibration displacements can be precisely measured with non-contact displacement sensors. When two displacement sensors are mounted offset by 90 and the signals are measured simultaneously beginning at 0 Hz, data is acquired on the static shaft position (gap) and the dynamic shaft movement (orbit) within the bearing plane. By comparing these values with the bearing clearance, information can be obtained on formation of the lubrication gap, especially when starting up, and on the thickness and position of the narrowest lubrication gap. Longer measurements also make it possible to evaluate vector circles or displacement paths. Eddy current sensors are among the most sensitive shaft vibration sensors. The sensor must, however, have a sufficiently large visible area (at least 200%). When the shaft begins moving (or vibrating) in the µm range, an electrical signal is produced. Both the DC and AC components are evaluated as a function of the distance. Measuring and evaluating mechanical shaft movement (part 2) Mechanical and electrical runout If the measurement trace of the displacement sensors contains irregularities, notches, scoring, scratches, drilled holes or surface flaws, mechanical runout results. Runout measurements belong to the acceptance criteria for new machine components and are described in detail in the API standards, for example. If runout is larger than 10% of the permissible shaft vibrations, the measurement trace must undergo certain surface treatments. Eddy current sensors are additionally affected by electrically induced runout due to microstructure differences, residual stress and residual magnetism. For this reason, shaft traces are often demagnetized. In Measuring principle of eddy current sensors The working principle behind eddy current sensors see Fig. 1 is based on the fact that the coil in the sensor head generates an alternating magnetic field whose field lines emerge from the sensor plane, pass through the object and then close again. The measurement field (alternating magnetic field) generates eddy currents in the electrically conductive object, leading to a loss in joules. These eddy current losses in the object increase as the distance to the object decreases. On the input side of the sensor coil, the eddy current losses are reflected in a change in the complex input impedance, which is measured and evaluated. An output signal proportional to the distance is formed, such as 0... 10 V or 4... 20 ma. On turbo-generator sets, the eddy current sensors are usually mounted when the sets are manufactured. They are permanently integrated in the bearing and machine housing and the measuring signals can be collected via separate outputs of the protective systems. Eddy current measuring equipment is relatively expensive and is of limited use contrast, electrical runout is not an issue with optical and inductive sensors. AMPLITUDE VS. TIME 3 The runout can be 1 2 identified through measurements during slow 1 0 T1 rotation, and can be -1 stored and used for compensation if necessary. 0 45 Fig. 1 illustrates runout 1 1 for a geometric eggshaped shaft. The posi- 0 0.5 2 3 T1 tion of the sensors is irrelevant 0.5-1 here. VERTICAL TIMEBASE WAVEFORM 0 45 HORIZONTAL TIMEBASE WAVEFORM 4 5 90 135 180 225 270 315 360 4 Shaft 6 5 6 90 135 180 225 270 315 360 7 7 S 8 8 T2 T2 9 9 S = > S = < S = 0 + 0 + 0 Fig. 1: Measuring principle of an eddy current sensor and definition of direction of movement. T I M E B A S E HORIZONTAL INPUT VERTICAL INPUT 0 AMPLITUDE VS. AMPLITUDE 270 X 2 4 Y 0 +1 3 5 1-0.5 9 + 0.5 90 6 7-1 180 O R B I T Fig. 1: The runout acts independently of the measuring direction. Oil Housing when employed in a mobile or temporary manner. Inductive displacement sensors and, increasingly, optical sensors are lower in cost and more flexible. Any nonlinearities, such as in inductive displacement sensors, are evened out either by PRÜFTECHNIK equipment through interpolations or by materialspecific characteristic curves that are stored in the setup. Inductive displacement sensors can be used to measure machines that are not equipped with displacement sensors at the factory. 2

Measuring and evaluating shaft movement (part 3) Analysis of shaft displacements and bending lines Rotating shafts are subject to varying degrees of play. During standstill, horizontal shafts rest in the lower bearing housing due to the force of gravity. When the oil supply is switched on in rotors with journal bearings, the shaft floats up slightly. Depending on the position of the oil supply, a certain lateral change in position may result. When the rotor shaft is set into motion, the shaft wanders upward in the bearing shell in a manner dependent on the direction of rotation (Fig. 1) and Bearing shell Shaft Lubricant reaches a more central operating position. If the shaft is accelerated even further, it rises still higher due to its inertia. The resulting pressure point must not lie in the area of the lubricating oil supply. At an infinite rotational speed, the shaft would concentrate in the center. This displacement is also exhibited during ramp down, a behavior that can be measured via the DC value. Irregularities in the displacement diagram point to rubbing, bearing damage, impermissible bearing clearance, incor- rect alignment or sudden load changes. By also evaluating shaft displacement in the other rotor planes, it is possible to plot the static and kinetic bending lines of the rotor shaft in the housing. Fig. 3 shows the static and dynamic bending lines of a shaft with three bearings. These types of bending lines can be used to predict how shafts will be displaced after the addition of load. It can also be determined whether forces, additional bending moments and additional bearing loads act on the shaft during operation. n = 0 n = slow n = fast n = Fig. 1: The rotor shaft wanders in the bearing shell in an rpm-dependent manner. Fig. 3: The static (blue) bending line and two dynamic bending lines of a shaft with three bearings. Fig. 2: The displacement of the rotor shaft depends on the direction of rotation and the shaft then moves along its orbit. Measuring and evaluating shaft movement (part 4) Shaft vibration analyses Amplitude spectra time domain analyses phase spectra 8 Vibration excitation sources can be identified on the basis of frequency and order spectra. After all, every rotating shaft produces rotational excitations that lead to shaft vibrations with more or less large amplitudes. First, however, it should be checked whether the time waveforms of the shaft vibration are stable and harmonic. Then the amplitude spectra can be examined for further excitations or even for natural frequencies. This method will also identify atypical excitations. If several multiples are discovered in the amplitude spectra, dynamic time domain analyses become necessary. Thus, tarnishing can lead to superimposed additional movements of the shaft. Also, phase dependencies can be more easily identified by time-synchronous measurements. Shaft vibration analysis should take into account that the measurements taken by displacement sensors are directionally dependent. If the sensors are mounted at -45 and +45 instead of 0 and 90, significantly lower amplitudes result at sensor A and higher amplitudes at sensor B with the same orbit. This is illustrated in Fig. 1. The procedure requires at least a 2-channel vibration analyzer because it prevents vibrational changes from appearing to occur only at one sensor because the orbit is turning. Continued at the bottom of page 9

Measuring and evaluating shaft movement (part 5) Unfiltered and filtered orbits 9 When a rotating shaft is displaced, it vibrates around the shaft centerpoint. The traces of the shaft around the shaft centerpoint are known as orbits. Orbit analysis is used to plot both the natural (unfiltered) and order-filtered shaft orbits in the bearing plane. Sudden changes in the amplitudes and form of the orbits are an early indicator of disturbing influences. The rotational orbits are a special feature that also provides information on the current balancing and alignment condition of the shaft. Continued from page 8 Fig. 1: The installation location of the vibration accelerometer influences the amplitude height. The form of the orbit is influenced by multiple factors. The most important of these are the rotational speed and the mode of operation. When an isotropic machine is operated at a subcritical level and without constraints, the orbits are circular. In contrast, elliptical orbits indicate the presence of forces due to alignment error or uneven bearing stiffness in two directions. If the orbit starts turning, contact a specialist. Mounting of shaft vibration sensors at positions 0 and 90 and -45 and +45 S MAX Fault Imbalance (thermal, magnetic, massrelated) Alignment Oil whip Transverse crack Too small local clearance Rubbing (labyrinth seals) Sensor A S P P Cause Forces from - deposits - asymmetry Forces from - constraints - asymmetry - increased play Oil film instability (steam turbine) Fatigue, stress Pressure to one side Reduction in radial play S P P Frequency spectrum Sensor B Sensor A S P P Orbit S MAX Sensor B S P P Glossary of terms Did you know? Shaft movements are exhibited as shaft displacements and shaft vibrations, acting radially and/or axially. The units of shaft movement are mm, µm, mils, inch. 1 mm = 0.001 m = 0.04 inch. 1 inch= 25.4 mm. 1 µm = 0.04 mils. 1 mils = 25.4 µm. Journal bearings have considerably larger bearing clearance than roller bearings. Practitioners work with the assumption that radial bearing clearance is 2 of the shaft diameter. Journal bearings can be differentiated into radial and axial bearings. Oval clearance bearings, multi-surface plain bearings and tilting pad bearings are special designs. Orbits describe the kinetic trace of the shaft center point. The gap is the distance between the shaft vibration sensor and the rotating shaft, and it changes with shaft displacement. Oil whirl generates shaft vibrations at the 0.38th to 0.49th order and results in a high probability of metal-to-metal contact. Unfiltered orbits jump significantly. Oil whip is a type of oil whirl with exactly half the rotational frequency. The machine runs with double the bending critical speed and excites half the rotational speed. The oil whirl frequency remains constant as the rotational speed increases. Anisotropy results from varying degrees of stiffness in the bearing, shaft and/or housing and causes directionally dependent vibrations. It must be taken into account in shaft vibration analyses. Phase angle is the rotational movement from the point where the trigger pulse is output to the maximum vibration amplitude. Critical speed: The rotational speed at which the system becomes resonant. Balancing: The procedure of improving the mass distribution of a rotor so that the rotor does not generate period forces on its bearings during rotation and the orbits remain small. Bode plot: A graph that displays the transfer function of a dynamic system. This type of analysis, which is also known as frequency response, stems from the field of electrical engineering from the 1930s and is named after Hendrik Wade Bode. Nyquist plot: A graph in polar coordinates that contains multiple vectors (magnitude and phase). The vector tips are connected by a line and the parameterization is specified for each one. Waterfall diagram: A graph in which the results are plotted against the rotational speed, independent of the measuring time. Corbits are waterfall diagrams of orbits in which the rotor speed is plotted on the Z-axis.

Measuring and evaluating shaft movement (part 6) Orbits of rigid rotors in flexible bearings Unfiltered orbits permit certain conclusions to be drawn on the type of journal bearing and the lubrication in use. Journal bearings must absorb the load, deliver the necessary stiffness and damping, and control the rotor position. The most commonly used journal bearings are divided cylindrical radial bearings. Oil is supplied through holes, grooves or pockets. Some journal bearings may only be operated in one rotational direction or unipolar bearings. If there is a risk of vibration, the machine manufacturer takes steps to reduce this risk, usually by changing the bearing type (bearings are characterized by their gap geometry, diameter-towidth ratio and load angle). Initially, he installs oval clearance or double-wedge bearings. If this is insufficient, the next step is to employ bearings with three or four sliding surfaces or even radial tilting pad bearings, which are the most expensive. Fig. 1 shows different bearing designs, which each exhibit specific characteristics in unfiltered orbits. All bearing designs have in common that they are flexible, a fact that needs to be taken into account when analyzing shaft movements and shaft orbits. They dampen vibrations and influence the directional vibration behavior. A special vibration-related feature of high-speed journal bearings is the selfinduced vibration of the rotating shaft via the oil film. The laminar lubrication gap flow becomes unstable when exposed to these self-excited vibrations, leading to ring vortices and even to turbulence. In extreme cases this can even cause journal bearings to transport drive energy into bending vibrations, resulting in whirl and oil whip. In this case, measures must be implemented that increase outer damping, the natural bending frequency of the rotor or the bearing topography. Fig. 1: Several journal bearing designs a) Standard circular journal bearing shell b) Oval clearance bearing c) Journal bearing with offset oval clearance d) Multi-surface plain bearing with 4 sliding surfaces e) Tilting pad bearing with 5 tilting pads Measuring and evaluating shaft movement (part 7) Shaft vibrations in flexible rotors Rotors that are operated above the bending critical speed are referred to as flexible rotors. Depending on the machine setup, speed, balance condition and rotor-dynamic properties, different kinetic shaft traces and/or different orbits occur at the individual shaft crosssections. When the behavior of a shaft with three separate discs is depicted in a diagram, areas with multiple critical speeds and different vibration modes can be discerned. 10 Fig. 1: Bending lines of a rotor with three separate discs in the critical speed range. Fig. 2: The frequency spectra with broadband excitations in the respective natural frequency ranges.

Measuring and evaluating shaft movement (part 8) Run-up & coast-down analyses of flexibly mounted rotors The vibration behavior of high-speed turbomachinery can be very complex. Rotors and bearings often exhibit vibrations in different directions. In horizontal machinery, the horizontal stiffness is usually less than the vertical stiffness, which leads to anisotropic behavior with different natural frequencies. In particular, run-up and coast-down analyses O W 3839 for a description of the available evaluation methods. h) Shaft movement diagram Both shaft displacements and superimposed shaft vibrations can be depicted as spatial shaft movements or bending lines in multiple dimensions (page 8, Fig. 3). Other display methods Modal and operating deflection shape (ODS) analyses, classification methods, shaker excitation and sweeps are further Fig.1: For natural frequency analyses, sensors should preferrably be mounted in a radialhorizontal and radial-vertical direction. provide the specialist with a tool to obtain more information on discrete machine excitations and natural frequency excitations. By analyzing shaft and casing vibrations during run-up and coastdown, it is possible to differentiate between housing or structural natural vibrations and natural vibrations related to the rotor system. There are various methods of plotting the measurement results: a) Bode plot Bode plots are the simplest analytical method. The phase angle and usually the speed-filtered component of the vibration signal are plotted against the speed. A turning point in the phase angle together with an amplitude maximum is a reliable indicator of resonance. b) Nyquist plot In the Nyquist plot, the magnitude and phase are displayed in a single diagram, making it easier to identify resonances and couple-critical speeds. c) Waterfall diagram For the waterfall diagram, frequencies and order analyses are measured at different speeds and superimposed as a function of speed. d) Campbell diagram In a Campbell diagram, natural frequencies are plotted against the speed on the abscissa and compared to typical excitation frequencies. e) Spectrogram A spectrogram is a time-variable, color-coded display of the frequency distribution based on a short-term FFT. Sound spectrograms are a special form and were formerly called sonograms. f) Order spectrogram Order spectrograms are a new PRÜFTECHNIK capability. During runup or when under load, resampled spectra are continuously recorded and superimposed on each other on a timedependent basis. g) Corbit (cascade orbit) Orbits can change spatially during run-up analyses. Please see the VDI Fig. 2: Imbalanced rotor with flexible bearings. display methods that make the vibration behavior of flexible rotors with flexible bearings during run-up and coast-down easier to understand. The vibration technology department of the VDI is currently working on a new standard in this field. 11 Fig. 3: Orbits change in their form, amplitude and rotational direction when passing through the critical speed. n krit n crit

12 News Online monitoring of wear Counting and size classification of wear particles in lubricating oil circuits is a growing field in Condition Monitoring and supplements vibration-based methods. The WEARSCANNER from PRÜFTECHNIK is a compact sensor system that detects electrically conductive particles in the oil flowing through it, counts these and classifies them by size as per ISO 16232. Changes over time in the size distribution allow users to draw conclusions on the wear condition and damage development of the components. The WEARSCANNER particle distribution counter is a valuable Condition Monitoring tool not only in wind power plants. One black line is all it takes The laser-optical trigger and rotational speed sensor (VIB 6.631) recognizes markings on the shaft. A black line positioned at right angles to the direction of rotation is usually sufficient as a trigger mark even on highly reflective surfaces. After being adjusted to the marking, the non-contact sensor delivers speed values and trigger pulses from a distance of up to two meters. This makes it easier to install and safely operate the sensor. Dates Information on all trade fairs, seminars and other important events of the PRÜFTECHNIK Group can be found on our website at www.pruftechnik.com Measuring shaft movement For the measurement of shaft movement, PRÜFTECHNIK has a special displacement sensor and various connection adapters: 1. Inductive displacement sensor (VIB 6.640) Direct connection to VIBXPERT Large working range (3 15 mm) Linearization of the characteristic line in the device Works with a considerably smaller detection area than an eddy current sensor Includes a universal holder for axial and radial installation A suitable substitute for dial gauges 2. Connection adapter for existing protection monitoring systems For measurement at buffered signal outputs and at the keyphaser output (VIB 5.433, VIB 5.332) With overvoltage protection in compliance with interface conditions for VIBXPERT EX (VIB 5.433-X, VIB 5.332-X) 3. Connection adapter for existing powered displacement sensors (VIB 5.341) For example, the IN 085 from SCHENCK. In the lead six times over VIBXPERT II is not only valued by customers. The international press, too, has singled out the fast vibration analyzer with numerous awards: six respected trade publications see VIBXPERT II as being outstanding in terms of innovation, power, flexibility and usefulness. More at www.vibxpert.com A license to measure You can become a certified vibration expert according to ISO 18436 in the category I, II or III. Manufacturers and operators of machinery are seeking certified and skilled technicians in the field of vibration analysis and diagnosis. PRÜFTECHNIK has therefore expanded its seminar program and now offers a three-part ISO CAT training that includes a standards-compliant qualification program with certification. These three series of seminars build on each other and take four days each. Participants who pass the final examination become qualified certified vibration experts in the corresponding ISO category, either I, II or III. PRÜFTECHNIK is represented in the technical committees that implement the requirements of the above-mentioned ISO standard in the Germanspeaking regions. PRÜFTECHNIK Condition Monitoring GmbH 85737 Ismaning, Germany Tel: 089 99616-0 Fax: 089 99616-341 email: info@pruftechnik.com PRÜFTECHNIK Alignment Systems GmbH 85737 Ismaning, Germany Tel: 089 99616-0 Fax: 089 99616-100 email: info@pruftechnik.com www.pruftechnik.com