Diagnosing and Curing Global Ship Resonances


Michael Bahtiarian, Noise Control Engineering, Inc.


(Originally Presented at the Insitute of Environmental Science & Technology Conference, May 2000)




FIGURE 1: "Outboard Profile" of Typical Offshore Supply Vessel or OSV



TABLE 1: Propulsion System Forcing Frequencies


ID    Hz         Description


1     2.75       Propeller Shaft Rotation Rate (RR)   

2     5.5         2nd Harmonic of Propeller Shaft RR

3     8.25       3rd Harmonic of Propeller Shaft RR

4    11.0        Propeller Blade Rate (4 Blades)

5    15.0        Propulsion Diesel RR (900 rpm)

6    22.0        2nd Harmonic of Propeller Blade Rate


At this point, in the process, we have defined the vibrational frequency at 2.75 Hz and confirmed the low frequency sensation. Narrowband data similar to that shown in Figure 2 were taken at various locations along the length of the accommodations area. These locations are identified by frame number, where higher frame numbers represent further aft in the ship. Data taken along the length of the vessel at the frequency of maximum vibration (2.75 Hz) was plotted as a function of frame number, Figure 4.



FIGURE 4: Vibration Amplitude at 2.75 Hz vs. Location on the Ship (higher frame numbers indicate further aft)



The source of the vibration based on the 2.75 Hertz frequency is the propeller output shaft. The most interesting fact about Figure 4, is that the shafts are located far aft in the vessel, at frames 85 to 103. There are no other mechanical equipment forward that would produce vibration at this frequency. The only way for the amplitude of the vibration to increase as you move forward in the ship would be a resonant condition, and this initial set of data suggests that the forward section of the ship is resonating in cantilevered beam mode.


So far the initial testing has confirmed that a vibration condition exists at a frequency equal to the propeller shaft once per rev. The amplitude of the vibration increases as you move more closer to the bow and further from the source. The final fact that complicates the situation is that the vibration condition is not consistent. The condition appears to change with vessel loading and the reasons for this needs to be understood.



FIGURE 5: Graph of Approximate Vessel Mode Shape Vibration on OSV Main Deck at 2.75 Hz.




Since the subject vibration condition did not occur during the detailed test phase, the author was left with speculation. Three elements affect the generation of the excessive vibration. They are (1) the ship's natural frequency, (2) the shaft operating speed and (3) the ship cargo loading. Fixing this problem obviously involves modifying at least one of these three elements.


Modifying the ship's natural frequency is accomplished by stiffening the hull, adding pontoons to decrease the displacement or installing a dynamic absorber (a large mass on springs). All options are a significant and very expensive undertakings. Either decreasing or increasing the shaft operating speed to avoid coincidence with the natural frequency would eliminate the vibration. On a permanent basis, it requires changing the reduction ratio of the gearbox and possibly the propeller. This is less costly than structural modifications, but still a pricey solution.


Correcting the structural or mechanical deficiency caused when the vessel is fully loaded, may provide the most economical solution. Suggestions for the next steps were provided and were to be carried out by other marine consultants. These steps included shaft alignment, inspection of the gearbox output shaft bearings, stem tube bearings and rudder bearings, and deflection computation at these bearing points for the various loading conditions.


Of all the recommendations, alignment measurements were conducted by the gearbox distributor under various cargo loading conditions. The alignment measurements showed that the clearances in the stern tube journal bearing were too small when the vessel is fully loaded. The location of inboard tanks and deck cargo is such that a bending moment is applied to the bearing and with improper clearance a high vibration at the shaft rotation rate was produced. This condition produced the input force causing the ship’s natural frequency to resonate.


This condition is consistent with the decrease in the third harmonic of the shaft rate found in Figure 3. When the shaft was unloaded, the bearing clearance caused a misalignment condition that was alleviated with load. Repairing the bearing ended the severe vibrations.




In retrospect, the natural frequency condition was the symptom not the disease, and the cure involved repairing machinery, not de-tuning the ship. Given the great cost to change the natural frequency of a ship, the restraint in suggesting such modifications were prudent. The data presented shows that vibration theory applies to structures of all sizes, including an ocean-going vessel. This ship and all its complexities still behaves structurally as a free-free beam. The important point to consider is that even though vibration theory applies, it does not mean that vibration theory will provide the solution, and other factors must be kept in mind in reaching a solution.




Vorus, William S., Principles of Naval Architecture, Chapter 7 "Vibration", Published by the Society of Naval Architects and Marine Engineers, 1988.


Vibration Control in Ships, Published by Veritec, Hovik Norway, 1985