When a piping system is exposed to fluctuating pressure disturbances, or pulsations, it frequently responds by vibrating. The magnitude and nature of the piping system vibration are dependent upon the frequency and energy content of the excitation. Low- to moderate-level periodic excitation, such as the pressure pulsations from positive-displacement or constant-speed centrifugal pumps, will not ordinarily excite significant levels of response in the piping system as long as the excitation frequencies are well removed from the natural vibrating frequencies of the pipe. If the pulsation frequency of the disturbance coincides with the natural frequencies of the piping system, however, resonant vibration can occur. Resonant response normally results in vibratory amplitudes many times that which would occur if the disturbance did not coincide with the natural frequencies of the piping system. Broad-spectrum or random excitation of the type associated with cavitation, bubble collapse, and extreme pressure reductions can also lead to resonant vibration. This type of vibration is known as self-excited vibration. The piping system draws energy from the broad-spectrum excitation and responds by vibrating at its own fundamental or harmonic natural frequencies.
Resonant response, whether due to the effects of fixed frequency or random excitation, can lead to unacceptable piping system damage. Cyclical stress reversals associated with resonant vibration can result in short-term fatigue failures, which may occur after only a few hours or days of operation. Reduction in the cyclical stress levels and attendant failures can be accomplished through a number of approaches.
When the excitation is a constant-frequency disturbance, ‘‘decoupling’’ the vi-brating piping system from the source of excitation can often be accomplished bychanging the frequency of the excitation. If the disturbance comes from a positive-displacement pump, changing the running speed will change the disturbance frequency. If a centrifugal pump is involved, a change in running speed or in the number of vanes on the impeller may have a beneficial effect.
Changing the natural response frequencies of the piping system can also mitigate the effects of fixed-frequency vibration. Again, the objective is to decouple or ‘‘detune’’ the piping system relative to the disturbance. This can often be accomplished by adding supplemental bracing to the pipe or by breaking the system intosmaller segments by introducing flexible elements.
When the excitation is broad spectrum or random in nature, detuning by changing the piping system natural frequencies is usually not effective in solving the vibration problem. The modified piping system will continue to draw energy from the broad-spectrum excitation and will vibrate at its new natural frequencies. Mitigation of this class of problems usually requires the reduction of the energy level of the excitation or the ‘‘strengthening’’ of the piping system.
A large number of broad-spectrum vibration problems are the result of high-differential pressure reduction systems. Excitation (noise) reduction for these types of systems can often be accomplished by the use of low-noise cage-type pressure-reducing valves or multiple orifices arranged for staged pressure reduction. In other cases, where the excitation is the result of turbulence, geometric changes to the piping system to smooth out the flow or reduce average and local velocities can have a beneficial effect.
The objective of the strengthening process is the reduction of piping system stresses to a level where fatigue failures are substantially eliminated. Much can be accomplished by the elimination of stress concentrations through the removal of geometric discontinuities. Examples of these discontinuities include hanger lugs, insulation supports, and small pipe (vent, drain, test, etc.) connections. Additionally, where changes in section are required, they should be effected by gradual, smooth changes in contour and generous fillet radii.
Special attention should be paid to all in-line welding done on pipelines that are subject to fluid-induced vibratory loadings. The use of inert-gas root pass welding with filler metal addition is recommended. This technique reduces the potential for the formation of critical root defects, which can lead to crack initiation and propagation. Radiographic and ultrasonic examinations done in excess of the mini-mum requirements may prove cost-effective in identifying stress-intensifying volumetric defects. Where such examinations are planned, weld backing rings should not be used since they complicate the job of interpreting the examination results.
Finally, the use of pipe wall thickness in excess of that required for pressure integrity design alone has been found to be beneficial in mitigating the effects of fluid transient loads.
An experience was encountered which required the replacement of the main turbine bypass piping on a large nuclear unit because of multiple short-term acoustically induced vibration failures. The original system was made up of NPS 30 (DN750) and NPS 24 (DN 600) X ³⁄₈ in (10-mm) wall pipe. The original design of the piping wall thickness was based upon pressure-integrity considerations alone. Several pressure boundary failures were experienced at pipe support lugs, hanger clamps, and vent and drain connections after only a few hours of operation. The replacement material was NPS 30 (DN 750) X 1-in (25.4-mm) wall and NPS 24(DN 600) X 1¹⁄₄-in (31.5-mm) wall. In addition, special attention was paid to the elimination of non-axisymmetric discontinuities and minor welded attachments. Lugs were replaced by rings, as shown in Fig. B2.2, and unnecessary small pipe connections were removed.
The upgraded system has been operated extensively without experiencing any further pressure boundary failures.
Since systems subject to flow-induced vibratory loads usually see those loads over much of their service life, their design should be based upon sustained loading criteria with no increase permitted in allowable stress.