An area where the down ratchet has had particularly pernicious effect is main engine shafting. Between 1998 and 1999, there were at least eight stern tube bearing failures on brand new VLCC's. A stern tube bearing failure generally leaves a ship adrift and helpless. The only good thing about these particular failures is that they occurred so rapidly -- in two cases before the ship was delivered -- that the ships had not yet loaded oil. These failures are shown in Table 13.1.
TABLE 13.1
DHI51091998-08-17Railko CY160LS
HHI10891998-05-08Railko CY160LS
HHI10901998-07-03Railko CY160LS
DHI51201999-01-05Railko CY160LS
DHI51211999-01-28Railko CY160LS
HHI11641999-10-26Railko ????
SHI12411999-12-02Railko WA80H
SHI12411999-12-02Railko WA80H
It is a rare month in which we do not hear reports or rumors of yet another shafting problem on young VLCC's. Thanks to the ability of this industry to hide its problems -- a practice in which the Classification societies play an important role -- we can be sure that there have been many problems we haven't heard about. The repair yards tell us that the new VLCC's are showing up at their first docking with very rapid bearing weardown. In July, 2001, HHI 1090 was again out of service with major stern tube bearing problems. Something is badly wrong.

Since all the Table 13.1 failures involved composite rather than white metal bearings, the immediate reaction was that there was something wrong with the composite material. Several owners replaced a newer composite material with a composite material that had proven itself for over 25 years in hundreds of large tankers, and which in the opinion of most of the tanker industry including the authors was superior to white metal. Two of these proven bearings failed almost immediately. Are we to believe a material that almost always lasts 15 or more years in heavy duty at sea service is the cause of dual failures within a few hours of installation?

Attention then turned to alignment. This was a natural assumption. The Class approved alignment procedure used by the yards is very crude. Over the years, the yards had somehow received class permission to bore the stern tube at the block stage, then weld the stern tube block in place, align by piano wire with the ship still on blocks, and hope. In fact, until about 2001 when a new LR limit on misalignment cames into force, there were no truly concrete requirements with respect to alignment. It was suspected correctly that the yards were taking advantage of the rules and the self-aligning characteristics of the composite bearings to be very sloppy in alignment. However in late 1999, LR carefully aligned two shafts using modern strain gauging techniques with the ship afloat at maximum alongside draft. Both these bearings failed before the ship completed trials. Alignment may be lousy but it is clearly not the root cause.

The current "solution" is to use white metal bearings and high volume, forced lubrication in place of the traditional oil bath system. This is a dangerous work-around, not a solution. The repair yards are reporting rapid weardown in the white metal bearings, and in our opinion, it's only a matter of time before they begin failing. High pressure lubrication is an invitation to blown stern-tube seals, and more importantly forces the the crews to make an impossible choice. If a stern tube seal starts leaking on a ship -- and this happens all the time -- the crew's normal response is to reduce the pressure in the stern tube lube oil system to nearly the same as the external sea water pressure. In an old style oil bath lubricating system, this generally halts the leak with nil increase in the chance of a bearing failure. If the crews attempt this with the current forced lubrication system, they face a high risk of a disabling casualty. If they don't adjust the pressure down, they face the certainty of a large fine at the next port, and a very displeased employer. This is the down ratchet in action.

To develop a real solution, we must understand the real cause of the problem. As Table 13.2 shows, over the last twenty five years, shaft diameters have decreased by at least fifteen percent for the same torque. This is a product of both higher strength material and the down ratchet. Since shaft bending goes as diameter to the fourth power, the net effect is that bending within the stern tube bearing has increased by more than 75% for the same propeller weight. At the same time, propellers have become bigger and heavier due to the decrease in Main Engine RPM. .sp 0.05i
TABLE 13.2
Hellespont Embassy, 197645,000851,010Smooth turbine torque
VLCC 199944,6407682080% torque pulses
There are no class restrictions on bending within the bearing. However, there is one class that has the capability to study at least a part of this problem. Bureau Veritas has a program which has the ability to model bending within the bearing and determine the resulting pressure distribution and oil film thicknesses within the bearing. Figure 13.1 shows a typical result. Figure 13.1 is based on an alignment that was acceptable to class on the grounds that the misalignment at the aft end of the bearing was small (1.0e-4 radians) and the nominal pressure (bearing load versus overall bearing area) was a reasonably conservative 6.3 bar, well below the Class limit of 9.0 bar. But what counts is the distribution of the pressure within the bearing and the standard Class approved method simply cannot address that issue. In this case,the BV results show the pressure on the aft 10% of the bearing averages 140 bar, well over BV's recommended (and none too conservative) max of 100 bar. At this pressure the film thickness is a miniscule 31 microns.

And here finally we come to the reason why the composite bearings have failed immediately while the white metal bearings have taken longer. There is only one area where white metal is better than composite but in that area it is far better. That area is heat conductivity. The conductivity of white metal is over 30 times higher than that of the composite bearings. The composite bearing relies on the lubricating fluid to conduct away the 7.5 KW of heat generated in a VLCC stern tube bearing. But most of that heat is generated in the high pressure portion of the bearing where the film thickness is much too thin to do the job. The composite bearing burns out almost immediately. White metal has great deal more ability to conduct the heat away itself so there is no immediate burn out despite the thin film thickness. But that doesn't change the fact that the pressures are very high, in fact far above the yield point of the white metal, rapid weardown will occur, and premature failure is inevitable. We will see a lot of VLCC's dead in the water due to bearing failures. The only real question is: how many will drift ashore?

ABS and others have correctly argued that the shaft must be more flexible than the hull in way of the shaft. Otherwise we will generate unacceptably high reaction forces on the bearings as the hull deflection changes with loading condtions and wave forces. This is true. But it is also true we must reduce the shaft bending in the bearing to avoid high local pressure. The solution is obvious: thicker shafts and correspondingly stiffer aft bodies.

Hellespont decided to go with a 15% thicker shaft than Class requires. This brings the shaft diameter almost back to the standards of the 1970's and reduces bending in the bearing by over 70%. We also put more than 200 tons of extra steel in the aft hull structure. With this system we were able to get the max pressure according to BV down to 50 bar from 165 bar. It also allowed us to go to a two bearing system, to obtain the flexibility we needed with respect to bending moment and shear force at the Main Engine coupling. It is clear to us that the reason why the mid-70's V's and U's have had relatively good shaft performance despite the crude analytical procedures was the shaft diameters that were used. The industry has a choice:

Our view is that (B) will almost certainly lead to (A); in effect that we should do both.

However, the BV-style analysis is still lacking in at least three areas:
  1. We must do a much better job of estimating hull deflection in way of the shaft and main engine for a range of loading conditions, and the results should be fed into the shaft alignment calculation. For a typical modern VLCC, the difference in the shaft-line deflection in way of the intermediate bearing between loaded and ballast conditions is of the order of 5 mm. But under current rules, the actual deflection for a particular hull is not known. Unless we properly account for this deflection, it's nonsense to worry about aligning a shaft to +/- 0.1mm. Furthermore it may be possible to reduce this deflection materially by beefing up the engine room structure. We believe that proper shaft alignment design will require a stronger engine room structure. This is discussed in Section 3 which talks about the various reasons why we need a full hull finite element model. Proper shaft and bearing design is one of these reasons.
  2. We must do a much better job of incorporating the time varying nature of the loads on the stern tube bearing. Figure 13.2 shows the vertical moment on the shaft produced by the propeller as it revolves for the DSME ULCC. This is a four bladed propeller so the pattern repeats every 90 degrees. Figure 13.2 takes advantage of this by showing only one-quarter of a revolution. Figure 13.2 shows that at full load on average the propeller imposes a moment on the bearing which lifts the shaft at the aft end of the bearing. This is due to the center of thrust being on average above the shaft line. This uplift is often offered as a reason why we don't have to worry about high pressures in the aft end of the bearing, although never in our experience with any real back-up. (LR goes so far as recommending 1.0 to 2.0e-4 radians designed misalignment in the static case on the grounds that the uplift will then align the shaft better when operating.)

    Figure 13.2 shows that for this ship at deep draft the average uplift moment is about 300 kN-m. But much more importantly, Figure 13.2 shows that this moment varies from an uplift on aft end of bearing of 700 kN-m at 70 degrees to a downward moment of 100 kN-m (0 to 35 degrees). In the ballast case, the moment is always pressing down on the aft end of the bearing. In short, the uplift is not always taking pressure off the aft end of the bearing.

    The propeller also generates vertical and transverse forces which are almost always ignored. Figure 13.3 shows that these vertical forces are strongly downward and can be as high as 12 tons which corresponds to a downward moment on the aft end of the bearing of about 250 kN-m. But the point is that the time varying nature of the propeller forces imposes additional demands on the bearing which are currently ignored by the Rules as are transverse moments and forces which are roughly of the same size, even when the ship is going in a straight line. We have the capability of estimating these forces, and the results should be fed back into the shaft and bearing analysis. The simplistic assumption of a steady uplift which will always reduce pressure in the aft end of the bearing is almost certainly wrong.

  3. We must incorporate the heat dissipation issue into the analysis. Once we have the film thickness and bearing pressure along the shaft for a particular operating condition, it is a fairly straightforward problem to use Reynolds equation to estimate the heat generated within the film and the resulting oil temperature rise. This in turn will change the viscosity which can then be fed back into the preceding analysis as necessary. In fact, Lloyds has a program (Program 1811 Stern Bearing Misalignment Conditions) which performs these calculations. But this program is rarely used in anger and never, to our knowledge, as part of a systematic shaft design process.

Finally, we must do all this for a range of operating conditions: laden, ballast, aft peak tank full, aft peak tank empty, engine hot, engine cold, RPM at full speed, and RPM during maneuvering and lightering, and make sure that the shafting system can handle all these conditions. This implies running the analysis over and over, but that's what the computer is good at. A particularly difficult case is lightering in warm sea water. The fact that a VLCC has to operate for extended periods of time at 10 to 15 RPM during lightering, an extremely tough situation for shaft lubrication, is completely ignored by the Rules. But the best alignment will be a compromise. An alignment that looks very good in one operating condition can easily be horrible in other conditions. Unless you examine a full range of conditions, you will never know.