The vb3000 was used to determine the cause of a mysterious vibration found in a wastewater aeration system.

Vibration analysis specialist Commtest Instruments
(www.commtest.com) recently faced an interesting underwater challenge presented by wastewater aerator provider Aquasystems (www.aquaturbo.be), based in Halle, Belgium.

Aquasystems is an internationally recognized designer and manufacturer of high-performance wastewater aeration, mixing, and decanting systems. The company”s systems are precision designed for wastewater operations, which, as a vital public service, depend heavily on the reliability and optimal performance of plant equipment. Left unchecked, the slightest vibration in an aerator rotor could cause rapid damage to seals and bearings and detrimentally affect reliability.

To ensure optimum operation, all new and modified product models undergo rigorous testing, and any faults are quickly addressed. Recently, however, test results of a modified aerator model had Aquasystems” engineers stumped for an answer, as vibration analysis of the model running above and underwater revealed two significantly different readings.

After spending a considerable amount of time trying unsuccessfully to determine and mitigate the cause of the vibration, Aquasystems came to the conclusion that it lacked adequate in-house equipment and resources and contracted Commtest partner company, Condition Monitoring Systems bvba, to help solve the problem.

Condition Monitoring Systems introduced Aquasystems” engineers to the Commtest vb3000 vibration analyzer. The aerator model was tested for vibration using the vb3000 alongside complementary measurement techniques (including waveform analysis, phase measurements, and bump tests). The results were outstanding and accurately pinpointed the cause of the vibration.

By connecting the vb3000 to waterproof sensors and an underwater optical tachometer, a clear imbalance of harmonic readings was detected at specific running speeds. Once the readings were plotted by Commtest”s Ascent software, it quickly became evident that resonance from the structural bending of the pump was the root cause of the problem.

Whole Lot of Shaking
In certain operating conditions, or in case of design changes, vibration problems can suddenly appear during the performance testing of Aquasystems” machines. Any high vibration levels can cause early damage to the seals and bearings and brings down the reliability of the aerator.

The turbo aerator considered here was a modified model that had a longer shaft for use in deeper basins. Initial tests of the machine running above water revealed no vibration problems, but when the machine was later tested underwater, the readings revealed a serious problem.

In earlier balancing attempts, Aquasystems was unaware the machine had a balancing problem, as only 1.25 mm/sec RMS was measured at running speed. But after measuring the aerator with an FFT analyzer, spectral analysis showed high radial frequency components of more then 20 mm/sec RMS at twice the running speed.

Determining the root cause was difficult, as it could have been the product of several conditions, such as hydraulic forces from the impeller, shaft alignment, coupling problems, torsional shaft vibrations, housing resonances, mounting resonance, and so on.

Various tests on the rig with and without rubber supports were performed in effort to reduce the vibration, yielding little or no difference.

Rooting Out the Cause
To investigate the failure mechanism on this machine, a vb3000 was used to gather more specific information about the vibration. Other measurement techniques included waveform analysis, runup/coast-down measurements, phase measurements, and bump testing. An additional runout test was necessary to further analyze the problem.

However, measuring the startup of the machine proved problematic. A normal startup should only take a fraction of a second and would be too fast to measure correctly. Also the runout of the machine was very short because of the large water resistance against the impeller.

Obtaining good runup measurements required installing a frequency controlled motor drive next to the test rig, then manually stepping up the machine speed from 1,000 RPM to 1,600 RPM in steps of 30 RPM.

With a Commtest vb3000 analyzer connected to an optical tachometer and an accelerometer attached radial on the motor, quasi steady-state amplitude and phase readings could be taken at every step.

After downloading the runup data from the vb3000 instrument to the

Commtest Ascent software, a cascade plot was drawn that showed the evolution of the second harmonic as a function of running speed. The resonance at 49.0 Hz stood out as twice the normal running speed of 1,475 RPM (or 24.5 Hz).

At first, the vibrations were low, but once approaching the running speed, the amplitude raised exponentially. A shaft alignment problem would have had the strong second harmonic from the beginning of the plot, because it is a force that is present at all speeds.

The first imbalance harmonic is visible at 24.5 Hz, but is very small (1.1 mm/sec) compared to the amplitude of the second harmonic of the running speed (20.5 mm/sec).

The Bode plot from the second harmonic (blade passing speed) was used to show amplitude and phase to visualize around the resonance frequency of 49 Hz. Its plot shows the maximum amplitude at 49 Hz and a 180-degree phase shift.

Thanks to the speed controller, the aerator line frequency could be forced to run at 52 Hz instead of the five Hz standard line frequency. The blade frequency passed completely through the resonance, and at 52 Hz the two-times running speed amplitude was already dropped at approximately eight mm/sec.

Problem Solved
The next step in the process was to define an adequate structural modification that would shift the resonance away from the exciting frequency of two times running speed.

Based on the formula for the calculation of the resonance frequency of a linear undamped system, it was clear that an increase in rigidity would shift the resonance frequency upwards.

Aquasystems” engineers created a new design to stiffen the pump structure with extra reinforcements. Four long strips of steel plate were cut and welded perpendicularly on the tube connecting the upper motor support with the bottom bearing support. This modification made the structure much stiffer for the first bending mode shape. The aerator was then tested again, and this time the results were good. The amplitudes dropped to normal values (0.65 mm/sec RMS at one times RPM and 2.75 mm/sec RMS at two times RPM).

A structural bending resonance at 49 Hz, close to twice the running speed, was the cause of the vibration problem. The high hydraulic forces at blade pass frequency played the roll of agitator. Running at two times running speed, the blade excited the bending resonance of the aerator, causing the large vibration component.

The cascade runup plot and the Bode plot demonstrated that resonance was the cause of the high second harmonic. The bending resonance frequency was moved away from the strong excitation force frequency by adding stiffness to the aerator housing.

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