Usually, turbomachines show satisfactory vibration levels during test runs at the manufacturer’s shop. However, vibration levels after the commissioning could be higher than desirable levels. Too often, after several months of operation, vibration could show some increasing trends. This is the usual story of many turbomachines.
The lightly-loaded rotor assemblies in turbomachines could present some high vibrations. The reasons behind high vibrations of turbomachines are discussed.
The initial steps to diagnose high vibration problems of turbomachines are:
1- The unbalance should be checked.
2- The alignment should be checked. An acceptable shaft alignment should be achieved during all operation scenarios (more importantly, normal operating conditions and startup).
3- The vibration analysis should be carefully checked and verified by proper vibration tests and measurements.
Importance of bearing and rotordynamics study
The importance of proper bearing type and design should always be insisted. Tilting-pad bearings are appreciated nearly for any turbomachine applications; in addition to very high stability offered, the preload capability inherent in tilting-pad bearing designs (and the offset pivots) can result in the satisfactory oil film stiffness and damping even in lightly-loaded rotors of many turbomachines. This can result in reduced levels of shaft vibration and the satisfactory bearing life.
For a successful rotordynamics study, the entire turbomachine train should be analyzed, and all the supports and bearings should properly be introduced into the dynamic model. From theoretical point of view, the turbomachine system is usually a statically indeterminate system, in which the flexibility of bearings/supports and the flexibility of shafts, will determine the load distributions. In other words, loads could be function of the stiffness of bearing and support which could vary with the speed and many other operational parameters.
Another important aspect is the “anisotropy” in turbomachine rotordynamics behavior. As a result of the anisotropy, there are coupled “pairs” of natural frequencies, for vertical and horizontal directions (closely related vertical and horizontal natural frequencies). For example, in a turbo-compressor, because of anisotropic effects, a pair of first natural frequency, “1.02×f1” for the first vertical natural frequency and “0.98×f1” for the first horizontal natural frequency, is obtained instead of the first natural frequency “f1” calculated by the isotropic assumptions. Sometimes an anisotropy effect may affect the second natural frequency even more. In the above-mentioned example, a pair of closely related frequency, “1.04×f2” for the second vertical natural frequency and “0.96×f2“ for the second horizontal natural frequency, is obtained (the second natural frequency “f2” calculated by isotropic assumptions).
Effects of faults on vibration
Many problems with turbomachines are related to the bearings and seals. High vibrations and faults could be caused by the balance shift because of the rotating component fits (such as impellers, blade assemblies, others), improper installation of couplings, misalignment, poor foundations, fluid related instabilities (such as an incipient surge, rotating stall, and others), the pressure pulsation reflection in piping sections, light rubs, seal rubs, the poor lubrication of the bearings, and so forth. These malfunctions could affect the vibration of turbomachines.
The vibration monitoring is important in determining the condition of turbomachines. For example, a high vibration may be caused from an excitation mechanism of a gas labyrinth seal or from an aerodynamic excitation caused by pressure distortion in a high-pressure turbo-compressor. As another example for a turbo-compressor, the dynamic excitation mechanism could be the gas labyrinth seal of a balance drum. A solution could be to provide a purge flow at the balance drum’s entry so as to block and prevent the forward swirling flow from the last-stage impeller leakage to the balance drum.
Coupling angular stiffness
The importance of coupling angular stiffness is usually overlooked for the rotordynmaics and the dynamic and vibration studies. It is rarely considered in a turbomachine design and the train alignment. The angular stiffness of some couplings (for instance, some metallic disk couplings) could be extremely high. This angular stiffness should be properly modeled in turbomachine rotordynamics; otherwise this can make the rotordynamics results invalid. Sometimes, to achieve a satisfactory rotor response and dynamic behavior, an angular stiffness limit should be specified for a coupling. In some cases, a coupling should be changed in an existing turbomachine to achieve an acceptable angular stiffness.
A high angular stiffness of a coupling could be a problem particularly for a turbomachine with light radial loadings on bearings. This issue could cause more problems on the overhung style turbomachines. Usually, a coupling with lower angular stiffness can result in a better mechanical separation of different rotors in a turbomachine train.
Thermal movements of one or more rotor supports should be limited to a reasonable range. Sometimes a misalignment at ambient-temperature is provided to compensate changes at operating conditions; this method is not usually recommended; however it has been used for many machines. In this case, an excessive ambient-temperature misalignment (in an anticipation of the effect of thermal movement) can exceed the available running clearance in bearings at the ambient-temperature condition; this could lead to the bearing damage as soon as the machinery is operated (reaching the operating speed) before any thermal expansion takes place.
There is a concern that a misalignment (or an alignment deterioration during operation) with a coupling that has high angular stiffness could lead to a high loading on turbomachine bearing(s). A high misalignment set at the ambient condition (a pre-set at the ambient condition to achieve a good alignment during the operating conditions) combined with a high angular stiffness of the coupling could result in extremely high load during the ambient start-up conditions; such a high load could result in an immediate damage to bearing(s), even if the shafts eventually go into proper alignment at normal operating temperatures. When using a high angular stiffness coupling, the driver and driven equipment will not behaving as two independent machines; instead, there could be strong interactions between machineries connected together in a machinery train.
The case study is about a high vibration incident in a low-pressure (LP) overhung-impeller compressor. After one month of operation, it was reported that the vibration amplitude of the shaft, particularly on the inboard bearing (the bearing near the coupling) passed alarm limits and approached to a trip set point. The compressor train had tripped because of high vibrations. Inspections after the trip showed some seal and vibration damages. The reason of the failure was an improper coupling selection.
A high angular stiffness coupling (a strong metallic disk coupling) was initially used which resulted in a very high loading on the bearing and seal combined with a misalignment; all these caused damages to these sensitive components. To solve the issue, the following corrective actions were taken simultaneously:
1- A suitable coupling with a lower angular stiffness was selected (a sophisticated diaphragm-type coupling).
2- A comprehensive rotordynamics analysis, considering coupling effects, was performed. This study confirmed the suitability of new coupling and predicted low vibrations.
3- Proper alignment limits were specified and controlled.
By these modifications, the compressor showed a satisfactory operation and a low operating vibration.
Amin Almasi is a Chartered Professional Engineer in Australia and U.K. (M.Sc. and B.Sc. in mechanical engineering). He is a senior consultant specializing in rotating equipment, condition monitoring and reliability.