session no. 6 — paper #1: comments and discussion — part ii

ASNE DAY 1980 TECHNICAL PAPERS

SESJION NO. 6 - PAPER #I: COMMENTS AND DISCUSSION - PART 11

INTRODUCTORY REVIEW

T H E BASIC PAPER BY Mr. Lyssimachos Vassilopoulos and Cdr. Fred M. Hamilton, USCG, “Longitudinal Stiffness Analyses for the Propulsion Shafting Systems of the POLAR Class Icebreakers,” was published in the Naval Engineers Jour- nal, Vol. 92, No. 2 (April 1980), beginning at PAGE 179.

During the first and second seasons of its deployment in the Antarctic region, the new U.S. COAST GUARD Icebreaker Polar Stor experienced longitudinal vibration problems in the center and wing shaft systems while breaking ice. To minimize and hopefully eliminate the vibratory responses, the COAST GUARD initiated several investigations some of which were discussed in the basic paper.

One of the items under study concerned the axial stiffness of the shafting systems that is contributed by the thrust bearings, their housings, and the foundations upon which they rest. In all, three sets of thrust bearing foundations were analyzed in great detail. Analyses included the original (as built) foundations, a modified configuration subsequently adopted for the wing shafts of the Polar Star and Polar Sea. and a reinforced configuration representing an upper bound on the foundation stiffness that can be achieved without drastic changes in the shafting system architecture.

In most of the paper the Authors dealt with the development of the three finite element models that were exercised with the NASTRAN Computer Program to predict the stiffness of the three foundation designs. The results of all the calculations were reviewed in light of the broader objectives of the overall program, and recommendations were given for further investigations that will aid in the understanding of the Icebreaker vibration problems.

COMMENTS AND DISCUSSIONS

By: Mr. Norman 0. Hammer Office of Ship Construction Maritime Administration

The Authors have prepared an excellent technical paper discussing a longitudinal vibration problem of great interest from at least two viewpoints. The first viewpoint is that of a shipowner who has discovered that his new ship will not function as originally intended. The second viewpoint is that of an engineer trying to ex mine the problem rationally, to analyze the available data, and then to propose some corrective steps that should be taken to remedy the problem.

Based upon my past experience, such problems are usually intertwined with contract administration problems between the designer, shipbuilder, machinery manufacturers, and others who have been involved in the design or construction of the ship. Therefore, perhaps one of the more interesting aspects of this specific problem may be what has taken place behind the

100

9

Naval Engineers Journal, October 1980

scenes as opposed to that which has been revealed in the basic paper. Not knowing the details, however, this Discusser must limit his comments to several technical points.

The first comment must pertain to whether or not there is a problem. For example, Figure 3 of the basic paper indicates that the maximum longitudinal vibratory centerline shaft thrust bearing movement is approximately 18 mils (double amplitude). Is this excessive? Based upon Merchant Ship practice, where longitudinal vibration levels have reached 34 mils (D.A.) and greater and the target level after correction was 20 mils (D.A.), the 18 mils (D.A.) on the Polar Star does not look so bad. But this is where the plot thickens. The Authors correctly point out that when the Polar Star operates in ice, substantially higher vibration levels are encountered. The exact magnitude has not, however, been measured.

While there is no direct counterpart in Merchant Ship practice related to ice loadings, it is known, however, that ships operating in shallow water and with large rudder movement will experience levels substantially magnified above those measured during the free-route conditions corresponding to SNAME Code C-1 practice. Also, another important factor in deteriora- tion of equipment is the operating profile, i.e., where a Mer- chant Ship operates at maximum horsepower over the 24-year life of the vessel, the Polar Star will be exposed to high levels only during relatively short periods. Therefore, a key question that must be answered is: What target levelshould the Authors be aiming ar achieving?

Assuming one has established the target level, the next task becomes that of exploring the various alternatives for modifying the ship. Here, this Discusser must take issue with a few comments contained in the INTRODUCTION. For example: “If vibration exists in service, two things can be done. First, the excitation magnitude, its frequency, or both can be changed. The magnitude can be changed by altering the propeller geometry or improving the inflow conditions. The frequency can be changed by altering the number of blades and/or shaft speed. Second, the system properties can be changed by altering its natural frequency or damping ratio . . . .”

The above passage represents the text book solution to solving problems. As 1 am sure the Authors have discovered by now, these solutions are not so simple to achieve once the cost of such changes have been estimated. Also, given the present state-of-the-art, each option also involves a substantial element of risk in that the desired objective may not be reached even if the most promising path is followed.

I would like to conclude my discussion by offering some comments on thefive recommendations outlined at the end of the paper. My comments are as follows:

1) Recommendations 1, 3, and 5 pertain to further analytical studies. While there is no harm to performing such work, the only true test of such analyses will be full-scale work - i.e., modification of the ship and measurement of the “before” and “after” conditions.

SESSION NO. 6 - PAPER #1 ASNE DAY COMMENTS

2) Recommendation 2 pertains to determining the longitudinal natural frequencies of both center and wing shaft systems. Again, the only valid answers are those determined by full-scale measurements.

3) Recommendation 4 mentions alterations may be required to the shaft, thrust bearing and foundation stiffnesses. These solutions limit the proposed changes solely to the “response” side of the system with no consideration given to altering the “excitation force” part of the overall system. While this strategy follows the least cost path, experience also shows that it will be very difficult to achieve a 50% reduction in vibratory movement. Therefore, if the target value is less than 9 mils (D.A.), the Authors probably will have to consider changes that will reduce the excitation forces.

Again, congratulations to the Authors on an excellent paper.

By: Cdr. William M . Simpson. Jr.. USCG Assistant Chief. Design Branch Naval Engineering Division U.S. Coast Guard Headquarters

I would like to thank the Authors for the time and effort they have devoted in documenting the subject matter and for the boldness with which they have asked for comments on this subject which is of great interest to the U.S. COAST GUARD.

I would like to offer a few brief comments on the calculation of the stiffness of the oil film in the thrust bearing and its effect on the overall thrust bearing stiffness. Based upon the analysis of tilting pad thrust bearings given in Reference [ 11 of this comment, the relationship between the oil film thickness and thrust for the POLAR Class thrust bearing is calculated as:

2 P

T = 3.21 N/h

where: T = thrust in pounds N = shaft revolutions per second h = oil film thickness at the pivot point in inches P

Thus it is clear that T = f (N, hp). The spring constant of the oil film, “k,” is then given by:

k = dT/dhp = a T / a h p + ( a T / a N ) (dN/dhp)

The first of these two terms is rather straightforward, but not so for the second term. The partial derivative of “T” with respect to “N” is easy, but the multiplier requires knowledge of the relationship between the longitudini.1 effect of ice impacts and the associated change in shaft speed. It is felt that this term should not be neglected as an ice impact which can cause the thrust to go to zero must certainly affect the shaft speed.

The average value of the thrust is also functionally related to the effective spring constant as can be seen by looking at:

3 P

a T / a hp = -6.43 N/h

It is easily seen that as “T” goes down, h increases, and therefore, the spring constant for the oil film goes down. Thus, the effective stiffness of the thrust bearing will be a function of the thrust being produced by the propeller at the time of an ice impact.

Looking at the thrust bearing stiffness while 200,000 Ibs of thrust is being produced at a shaft speed of 175 rpm, and assuming a 10% reduction in shaft speed due to a 300,000 Ib ice impact (to calculate dN/dhp), it is estimated that the thrust

P

bearing stiffness would be 23% less than the Authors have calculated.

I look forward to the Authors’ reply to my comments, and I also would like to add my endorsement of the Authors’ recommendation that the non-linear aspects of the shafting system be the subject of future study.

REFERENCE

[ l ] Shaw, E., et al., “The Performance of Tilting Pad Thrust Bearings”, Tribology Convention 1971, The Institution of Mechanical Engineers, London, Eng., 1971.

By: Mr. John R. Kane Consultant Newport News, Va.

Icebreaking in the Polar regions is no business for timid souls, and one must admire the boldness with which the U.S. COAST GUARD approaches it, particularly in the development of powerful Icebreakers such as the Polar Star. Considering the complexity of the CODOG plant with which the ship is powered, and the tremendous amount of power packed into a hull of that size, it is not too surprising that the ship experienced vibration problems in her propulsion shafting and machinery while breaking heavy ice. This Discusser, having been present on the trials of several NAVY capital ships in the early 1940s which experienced severe longitudinal vibration, can appreciate the Authors’ statement that “the ship’s machinery and her complement were subject to an awesome level of vibratory energy” during the most severe icebreaking conditions.

While no one wishes trouble for such pioneering projects, there is a brighter side to it when it occurs, because only by get- ting into some degree of trouble can the engineering profession learn how to stay out of it. For this reason, it is to be hoped that the research reported in this paper is only the start in reporting on the various aspects of the vibration experienced, and that a more complete investigation and report on the vibration will be forthcoming in the near future, as was done on the longitudinal vibration problems in the Navy’s Battleships and Aircraft Car- riers in 1949 (see Reference [2 ] of the paper).

Several aspects of the longitudinal shafting systems of the Polar Star strike this Discusser as being unusual.

First, in order to keep a conservative factor of safety in the propulsion shafting systems with greatly augmented thrust and horsepower for icebreaking operations, the propellers and shafting are substantially more massive and rigid than would normally be the case with a ship of this size. This results in the situation that the foundation spring that takes the thrust bearing loading into the hull has a lesser spring constant than any of the shaft spring constants; a most unusual occurrence in ship shafting systems. In other words, the thrust bearing foundation was not beefed up in the same proportions as the shafting and propellers in order to take up the increased load due to icebreaking without excessive deflection. The correction of this oversight is the principal thesis of the paper.

Second, in order to make possible the separate gas turbine and diesel modes of propulsion in the CODOG plant, and in particular, the diesel electric propulsion mode without the gas turbine booster used for heavy icebreaking, the main shafting just abaft the reduction gear and forward of the thrust bearing contains a splined or dental disconnect coupling which, even when engaged, serves to sever the longitudinal vibratory system abaft of that point from the machinery and shafts forward of said point except to the extent that friction or “sticktion” of the coupling transmits the disturbance forward. When disengaged,

101 Naval Engineers Journal. October 1980

ASNE DAY COMMENTS SESSION NO. 6 - PAPER #1

this coupling simplifies the axial vibratory system, but when engaged it complicates the situation because friction in the coupling teeth can be so variable (This friction is of the “stick- slip” type and the starting friction varies tremendously with the material hardness and finish of the coupling. the lubrication or lack of it, the torque, and the steadiness and amplitude of the vibratory motions transmitted to it). A similar situation to this was described in System B of Reference (21 of the basic paper, where the thrust bearing was replaced some distance aft in the lineshaft and the response of the system was found to be highly dependent upon the character of the friction in the couplings and of the vibratory amplitude at the coupling as compared to the amplitude at the propeller. Unfortunately, not much ap. pears to. have been added to that knowledge as to the behavior of such couplings beyond that which was described at some length therein.

Third, the disturbing force produced by milling through ice under severe conditions with the propellers most likely doesn’t fit the usual assumption of steady state damped forced vibration produced by simple harmonic forcing functions at all well since it undoubtedly comes closer to a repetitive series of discontinuous pulse trains than to a simple harmonic oscillator, and free vibrations may enter into the picture in an important manner. On the other hand, the very unevenness and the impulsive or impact nature of such ice forces, even though they may be buffered substantially by the propulsion shafting system, will probably serve to keep the disconnect coupling alive and sliding when engaged, making the vibratory critical analysis more determinate in the gas turbine icebreaking mode.

Icebreakers, of course., must operate a t substantially augmented thrust in combination with reduced ship speed while doing their job; conditions well known to bring out double blade rate frequency, and sometimes even triple blade rate, more prominently than in free route operation. With the controllable pitch propellers on two different shaft systems, and considering propeller blade rate frequency excitation, 2 x blade rate and 3 x blade rate, and the several modes of propulsion, there are a remarkable number of shaft critical speeds in the possible operating range and at significant power for this ship. There are, for example, sixteen such criticals of the torsional type, eight of which involve the propeller as an antinode and thus likely as a source of excitation. Considering only 1st and 2nd mode torsional criticals, there are eleven in the significant operating range above about 60 rpm.

As for longitudinal critical speeds, a brief look-see by this Discusser indicates that there arefive corresponding first mode criticals in the operating range, and one slightly above the operating range, but which would have significant buildup at maximum rpm. When one thinks of trying to operate clear of all these potentially bothersome criticals, torsional and axial, with the gravelly propeller input one might expect from some tough old multiyear ice, it gives one cause to wonder. In addition, the possibility of coupling between axial and torsional vibrations through the medium of the propeller becomes very real. In fact, from some rough figures made by this Discusser, it appears that in at least one case, i.e., the DE propulsion mode with the disconnect coupling disengaged. the first mode vibratory torsional natural frequencies of both the wing shafts and the centerline shaft coincide within one or two rpm with the corresponding first mode axial criticals of those shafts. If this is correct, such coupling probably does exist, and, even if it doesn’t, one should be concerned by the maximum shaft stresses resulting from such superposition.

Turning to the stiffness analyses discussed in the paper, and to the thrust bearing stiffness in particular. this Discusser would point to the data presented in Reference (21 of the basic paper concerning thrust bearing stiffness in which a correlation

102 Naval Engineers Journal, October 1980

was obtained between theoretical calculation and a bench test of the bearing made on a large materials testing machine. The data in Reference [2] indicates that the main thrust bearing on the Polar Star may in fact be a bit stiffer than arrived at in the paper. It is not clear in the paper whether it was taken into account that the largest single component of deflection shown there, i.e., the spherical indentation between the shoe supports and levelling plates, is proportional to the two-thirds power of the load rather than linear with the load. To get a correct equivalent linear spring constant then, one must take the slope of the Deflection Curve in the vicinity of the mean load involved rather than basing the constant on the total deflection at that point. This results in reducing the influence of the total nonlinear spherical indentation somewhat.

With regards to the estimates of foundation stiffness, the ap- plication of the Digital Electronic Computer Program is most interesting. At the time Reference [2] was being prepared, the digital computer was still in its infancy, and the only computer service available to the Authors consisted of a couple of patient young ladies operating hand cranked Monroe calculators. Plex- iglass models were used extensively, with dial indicators to read deflections, and small vibrators were used to shake out the friction and to obtain repeatable readings - but only to get qualitative and relative results, never for quantitative numbers. The principal lesson learned was the importance of shear deflection in such structures, and that a good understanding of shear stress distribution and shear lag was more important than trying to bend the beam theory to fit the case. F.M. LEWIS, on the other hand, always saw stiffness in terms of a maze of can- tilever beams. Nowadays you can set up a finite element program on a third generation computer and let it crunch out the answer for you, and the results are probably more accurate. In any case, the value of about 11 X loL Ibs per inch for the wing shaft foundations seems quite reasonable, and the conclusion that it would be hard to raise the spring constant for the foundation higher than 25 X lo6 Ibs per inch probably quite correct.

By: Capt. L. C. Melberg. Jr., USCG (Ret.) Maintenance Director Washington State Ferries

The key result of the study of the shafting system seems to be the same critical question. Knowing what the present study has confirmed, what are the owner’s plans to resolve the longitudinal shaft vibration? The presently uncontrolled forces will continue to cause significant progressive damage on each ice-breaking mission and eventually lead to main motor failures (I would suspect soldered commutator, electrical joints, and excessive bearing wear); main reduction gear failures; and a con- tinued myriad of auxiliary failures such as in piping.

The present study very adequately has defined the problem boundaries in some areas and has begun to define some other areas. The Authors are to be congratulated on their problem solving methods.

Some historical background to the solution also appears in order. In the early stages of the design, studies concluded that the structural design would be a transversely framed grillage system which would optimize normal strength requirements for ice-breaking. The design work was checked using the same NASTRAN Program. Another in-depth study was also made in vibration, as was reported in Reference ( I ] of this comment. The structural studies did not include any detail foundation designs, but the vibration work correctly identified the critical


- 160 ln 2 150 > V 140 W

130 E

2 110 2 2 I2O

100

90

80 8 10 12 I4 16 IL) x ) 22 24 26 28 30 32 W

FOUNDATION STIFFNESS ( Id LBYlN )

Figure 1. [I]

nature of the longitudinal vibration. At that time, the scope and magnitude were presented in the study with nominal values for thrust bearing stiffness of a range of 7 X lo6 to 11 X lo6 Ibf/in and foundation stiffness of 10 to 34 X lo6 Ibs/in. Criticals were noted in the 170 to 240 rpm range as shown in Figure 1, which was reproduced from Figure 53, page 118, Reference 111.

Another important set of factors also turned up in the propulsion studies. When studying the electrical propulsion control, a model was developed for ice flow into the propellers. The results were presented by LEWIS in Reference [2]. He showed some interesting relationships between current ice torque and propeller speeds. The propeller forces due to ice contact are broad-band impacts, and were shown to have considerable influence on the control design. My purpose in referencing these two preliminary design studies is to recognize that the realprob- lem was not fully understood in preliminary or contract design and certainly was not by the builder.

We have all assumed that an Icebreaker inherently is a very stiff ship, and is structurally capable of handling any propeller loads, thrust bearing reactions, et cetera. As was discovered in preliminary design, the longitudinal strength was still important (controlling, not ice strength) in some areas of design (upper decks). It is not obvious that the transverse frame and grillage structure of the POLAR Class does not provide a SUE- ciently stiff structure longitudinally to handle the thrust and propeller ice-breaking forces. A look at the deflection diagrams presented here reveals the lack of longitudinal strength. Because of the broad frequency range of the ice propeller forces and the weakness in the present thrust bearing, if we are to continue use of the high powers, a very costly change must take place. The stern’area of the ship will need to be re-designed to include significant longitudinal strength (stiffness), and new thrust bearings which have some form of damping must be sought. One might even look for a viscous damping device to be installed around the shaft in the stern bossing.

REFERENCES

[ l ] “Integrated Vibration Study for New Design Polar Ice- breaker,” NKF Professional Engineers (10 November 1967).

[2] Lewis, J., “Some Aspects of the Design of Diesel Electric Icebreakers Propulsion Systems,” Naval Engineers Journal Vol. 81, No. 2 (April 1969) p. 90.

By: Lt. Robert W. Gulick, USCG Chief. Icebreaker Section Maintenance Branch Naval Engineering Division U.S. Coast Guard Headquarters

The Authors are to be commended for their lucid and in- structive presentation of this complex subject.

This discussion will focus primarily on a macro view of the COAST GUARD’S approach to this problem of which the Authors’ efforts represent a significant part. The following items provide the history and plans of the attempt to resolve the POLAR Class longitudinal vibration problem:

1) Identification of unacceptable longitudinal shaft thrust bearing housing displacement during DEEP FREEZE ‘78.

2) Investigation indicated severe longitudinal shaft vibration resulting from low system stiffness, particularly in the foundation, and a wide-band, high-energy excitation forcing function. 3) Determination of a modification subject to the following

constraints: a) a repair philosophy that dictated that improvements to an unacceptable problem area be made prior to upcoming deployments; b) operational commitments; c) availability of funds; and d) design lead time.

4) Identification and installation of interim fix (significant contribution made by the efforts described in the basic paper).

5 ) Installation of full-scale instrumentation package to assess the effectiveness of the modification.

6) Commencement of concurrent long and short term analytical studies by DTNSRDC and MARITECH, Inc.

7) Evaluation of full-scale data and validity of FEM studies. 8) Determination of follow-on efforts.

Completed analytical efforts and full-scale data collected during DEEP FREEZE ’79 indicate that returns from improvements to thrust bearing foundations stiffness are small. Accordingly, the major efforts will be redirected to mitigate the forcing function. The avenue of pursuit will be to bring the ice deflector project to fruition with an expedited installation in late summer 1980 on Polar Star. The project is now in the final ice model testing phase. Concurrently, a long-range effort to broaden the bas$ understanding of ship system vibrations will be con- ducted.

Propeller/ice-induced ship vibration is an old and continuing marine problem. The POLAR Class experience has confirmed two adages: it is very costly to implement a solution in retrofit instead of design mode, and vibration is principally a ship system problem.

The Authors are again commended for a well constructed and meaningful presentation of this complex area of common interest to the marine field.

By: Mr. T.H. Doussan Vice President & Chief Engineer Avondale Shipyards. Inc.

It does appear that the thrust bearing foundation lacks SUE- cient stiffness. This stiffness should be in the order of 20 X lob I b h . We have also previously calculated the stiffness of thrust bearings to be in the order of 30 X lo6 Ib/in.

The basic paper deals extensively with applying the finite element method to the analysis. However, it does not compare calculations for the basic system to measurements on the ship - that is, the Authors’ Figure 5 plots the thrust bearing model in its undeformed and deformed conditions, and their Figure 6 does the same for the side girder and bottom structure. Their paper, however, does not state whether the deformations



calculated for the thrust bearing match those observed on the ship. Also, it does not comment on whether the thrust variation in ‘the ship was measured.

It would appear, from a review of Figures 5 and 6, that the basic problem is deformation of the thrust bearing support structure, which can be somewhat limited by the modification applied in APPENDIX A, Figure A-5. This modification essen- tially provides a support at the upper level of the foundation, thereby resulting in a reduction of the corresponding displacement.

Stiffnesses of pedestal-type foundations will normally be 10 to 13 X 10’ Ibflin. while those extended athwartships to the shell are expected to be 20 to 25 X lo6 Ibflin.

With the thrust bearing located adjacent to a bulkhead, the deformation shown for the bottom could be expected. Again without going through all of the calculations in detail, it would appear that the basic problem is the supporting structure for the thrust bearing foundation rather than the foundation itself. To add validity to the study, some comparison of measured

thrust variations and deformation is necessary.

By: Dr. R. Wereldsma Professor. Ship Structure Laboratory De@t Universiy of Technology. The Netherlands

With appreciation I want to address myself in writing to the Authors in connection with the stiffness analysis of the propeller shafting in longitudinal direction. Having understood the problem reasonably well, 1 believe that natural frequency resonance phenomena, although important for the operation under icebreaking conditions, will not be the essence of the problem. Due to the broad banded and impulsive propeller excitation during icebreaking, these resonance phenomena will occur any time as long as the natural frequency is within the excitation frequency band. Therefore, the solution must be found in an increased stiffness or artificial damping, and indeed a stiffness analysis is necessary in the first place.

As outlined in the basic paper the stiffness of the shafting consists of many elasticities. Including the bearing stiffness of approximately 20 X lo6 Ibf/in and the original foundation stiffness of 10 X lob Ibf/in., we have in total a stiffness of approximately 7 X lo6 Ibf/in. Depending upon the broadness of the excitation, this stiffness must be increased so that the longitudinal natural frequency of the propulsion system falls out of the range of excitation frequencies. It is questionable whether this increase can be obtained without making drastic alterations to the structure and bearings. I should be very pleased to have the Authors’ views on this.

Further, 1 should like to draw attention to the fundamental operation of the propeller, i.e., the conversion of torque into thrust. This conversion also exists mutually for the vibratory mode operation, and a longitudinally stiff operating propeller may generate strong thrust fluctuations caused by torsional vibration. In fact, for a proper analysis, the two modes of vibration need to be dealt with simultaneously by coupled equations [I]. Strong torsional vibrations may occur easily because the torsional support of the propeller is, generally speaking, very weak (natural frequency well below blade rate).

My final question is concerned with the effects of the oil film. The flexibility of the oil film amounts to one-tenth of the total flexibility of the bearing and its components (Authors’ TABLE 3). and even a smaller fraction, when the foundation flexibility is taken into account. Is it acceptable, in view of the total analysis, to neglect the thrust bearing oil film effects, or to have a simplified analysis applied for the lube-film, in contrast to the


transverse direction where it seems to be necessary to make a full analysis of the lube-film dynamics with perpendicular mutual couplings (horizontal-vertical)?

REFERENCE

111 Hylarides, S. and W. van Gent, “Hydrodynamic Reactions to Propeller Vibrations,” Conference on Operational As- pects of Propulsion Shafting Systems, Institute of Marine Engineers, London, Eng., March 1979.

By: Monsieur Guy C. Volcy Deputy Director, R&D Division Bureau Veritas Paris, France

First of all, I wish to express my thanks to both Authors for their invitation to contribute to their paper. Second, I seize this occasion to compliment them for a valuable and very interesting paper. Third, I wish to tell them that they are very lucky to meet such an interesting case in which the longitudinal vibrations of the propulsion shafting system have amplitudes of the order of f 6 . 0 mml

On the basis of my rather long shipboard experience I wish to confirm that for an installation of 20,000 SHP, amplitude of f0.3 mm would be considered as lying within limits that could be termed tolerable. I, therefore, agree with the Authors when they describe as “awesome” the vibratory levels encountered on the POLAR Class Icebreakers, and that such vibrations may result in casualties to auxiliary machinery, electrical equip ment, and even rupture of piping.

I can hardly imagine that two such important ships were not studied at the design stage from the point of view of vibration. 1 suppose they were studied, but unfortunately in the conventional, old-fashioned style, namely, by investigating hull girder vibrations and not looking for the presence of forced vibration resonators. In actuality, the elastic system comprising the line shafting, its foundations, the steel-work, and the masses of the double bottom in the Engine Room constitutes a forced vibration resonator, which, by its resonant response, is able to lead to awesome vibratory levels in the ship.

At the beginning of my career, I also used the conventional approach, but very shortly my shipboard experience proved to me that what is most important is to look for the presence of such vibration resonators, which after detection could be de- tuned [2]. Having abandoned the conventional approach in 1969, I have since always avoided such annoying vibratory phenomena by forbidding the increase of local vibratory levels in the vicinity of such resonators.

For a successful study of vibratory phenomena it is always important to adopt an integrated treatment of static and dynamic phenomena since these two are interdependent [4]. I am therefore very pleased that in this very interesting and rich-in- information paper, the Authors also look first for the correct assessment of the static phenomena (static stiffness of foundations and tilting) before going over into the vibratory problem (in which the dynamic stiffness is taken into account). In actual fact, such an integrated study was introduced in BUREAU VERITAS about ten years ago and applied to many ships at the project stage, even to the biggest Tankers in the world, those of the BATILLUS Class, [7].

1 also must inform the Authors that I am very pleased with their correct interpretation of the encountered phenomena as well as the appropriate and reasonable assessment of the size of the steelwork to be modeled. This is most difficult and expensive to determine when it concerns the evaluation of the stiffness of the thrust block foundations. I have some objections,


however, as regards the constraints applied to all grid points that lie on the center plane of symmetry. In the symmetric case, the rotation must be zero about the Y axis, not the Z axis. In the anti-symmetric case, the rotation must be zero about the Z axis, not the Y axis. I think this may be due to a typewriting er- ror, and I would like to have the Authors’ confirmation on this point.

In vibration problems there are s i r different parameters present, the mass, damping, stiffness and the intensity, frequency and phase of the excitation. It is always prudent before making any decisions to assess their mutual importance on the basis of correctly executed calculations. After reading the basic paper I see that the Authors have adopted this approach and they should therefore be complimented for the enormous amount of work involved in studying this troublesome case. I also agree with them when they express their opinion that the over- simplified model of their Figure 1 is not adequate. Such an oversimplification can be the origin of vibration troubles since the thrust bearing foundations and double bottom flexibility are not taken into account.

I also was impressed by the correct and laborious way of calculating the differing amounts of stiffness of the thrust bearing as described in the basic paper. In this respect, I would like to present some comments stemming from our experience.

We agree with the Authors that the oil film stiffness is normally larger than that of the collar, but by a factor of approximately 2 not 4. I can mention the example of the EL PASO Tanker of 125,000 m’, built in the Chantiers de France- Dunkerque, and equipped with a single screw absorbing 45,000 SHP 161. The oil film stiffness was calculated to be 36.6 X lo9 N/m. The global stiffness of the thrust bearing, exclusive of the oil film, can be compared to the stiffness of the Stal Laval thrust bearing which is 8.0 X lo9 N/m. In this respect, I draw attention to the fact that the Stal Laval Kingsbury Thrust Bear- ing has flanges situated at the axis of the line shafting, which is considered to be a rational design. From the Authors’ paper I could not, unfortunately, deduce which type of thrust bearing had been adopted on these Icebreakers and whether by chance it is an old-fashioned one in which the thrust bearing is fixed on its foundation by means of flanges situated below the thrust bearing. As regards the stiffness of the thrust bearing foundations of

the EL PASO Tanker (which is generally considered to be vibration-free), it can be said that from the beginning of the design effort, especial attention was paid to the rational configuration of the thrust bearing foundations. The stiffness of these foundations, being particularly robust, had the value of 4.71 X lo9 N/m. If we compare this value with the stiffness of the “as-built” POLAR Class foundations, namely 1.957 X lo9 N/m, it is obvious that the latter is very small and that only the stiffness computed for the “re-inforced” foundations (Model 3) may be compared with the EL PASO Tanker stiffness. Trials and investigations of this ship have shown amplitudes of longitudinal vibration of the order of +0.025 mm, but these were made in open sea, not under ice-breaking conditions. In this connection, 1 also wish to mention the case of the specially- conceived thrust bearing foundations for the biggest Sea-Barge Carriers in the world, those built by the Valmet shipyard 181. These ships are free of vibrations and their longitudinal stiffness in the foundations is 7.63 X lo9 N/m.

From what I have said so far, a general idea may be formed, and that is, the necessity for conceiving at the project stage the rational design of the thrust bearing itself as well as its foundations. In fact, the thrust bearing should be “held by the ears and not by its feet.” This leads to the abandonment of the old- fashioned thrust bearing with flanges situated well below the shafting axis. If this principle is not respected, then severe

vibratory phenomena can be encountered and also serious damage to the gearing, such as those reported in References [l], [3], and IS]. Last year, BUREAU VERITAS published a new Guidance Note [9] which underlines the importance of conceiving at the project stage the rational design of lineshaft foundations.

On the basis of the preceding experience, the modifications to the thrust bearing foundations envisaged by the Authors could be termed rational. A supplementary solution could also be recommended, namely, the fixing of the thrust bearing to the shell plating by means of struts. Such a solution has been adopted and tested, yielding positive results. Of course, like always, consideration must be given to the balance between cost of modification and the expected improvement. In this respect, the Authors are very fortunate as they will be able to correlate their calculations with the model tests carried out a t DTNSRDC. It would be hazardous however to extrapolate the model stiffness from that determined by the Authors for real ships. In this respect, it would be more rational to carry out a calculation of the stiffness of the model to be tested.

As regards the issue of symmetric and anti-symmetric modes, the latter are needed not only when thrust is reversed, but also for normal operations of multi-screw installations. From my past experience I also would like to draw the attention of the Authors to the prudence of drawing conclusions on the stiffness values of casings from which the thrust bearing is made (basic paper, Figure 19). Haw all calculations described in the paper been made for the tilted or untilted condition of the thrust bearing? Did the Authors evaluate the damping properties of the thrust bearing and the foundations? From our experience, this type of damping is usually very low.

Before ending my contribution, I would like to express my interest in seeing the results of further studies by the Authors and in particular the data analysis of the random excitations caused by the propellers during ice breaking, calculations of forced vibration responses of the line shafting and correlations with experimental data, and finally, calculations concerning the nonlinear vibratory behavior of the system.

Once again I offer many compliments for this very interesting and informative paper together with my best wishes for the con- duct of future investigations and the final solution of the existing problems.

111

121

131

141

151

1 71

I61

I81

REFERENCES

Volcy, G., “Actual Behaviour of Marine Gearing. and Alignment Conditions of Line Shafting”, Nouwautes Techniques Maritimes (1968). Bourceau, G. and G. Volcy, “Forced Vibration Resonators and Free Vibrations of the Hull,” Nouwautes Techniques Maritimes (1969). Volcy, G., “Damdges to Main Gearing Related to Shafting Alignment,” IMAS Conference, London, IME, 1%9. Volcy, G., H. Gamier and J.C. Masson, “Chain of Static and Vibratory Calculations of Propulsive Plants and Engine Room of Ships,” ATMA, Bulletin (1974). Volcy, G., “Reduction Gear Damages Due to External In- fluences,” New York Metropolitan Section, SNAME, 1974. Volcy, G., “Vibrations of Aft Parts of Ships,” PRADS, Tokyo, Japan, 1977. Volcy, G., M. Baudin, and P. Morel, “Integrated Treat- ment of Static and Vibratory Behaviour of Twin Screw 553,000 TDW Tankers,” Spring Meeting Paper, RINA, 1978. Volcy, G. et al.. “European Built Sea-Barge Carriers, Their Design, Machinery Hull Interaction and Investigation into



Vibratory Behaviour," Institute of Marine Engineers, Lon- don, Eng., 1979.

[9] Bureau Veritas, "Recommendations Designed to Limit the Effects of Vibrations on Board Ships", Guidance Note NI 138A. 1979.

By: Mr. F. Everett Reed President & Technical Director Littleton Research & Engineering Corp.

I am pelased to be able to present a discussion on this interesting paper. MR. VASSILOPOULOS and CDR. HAMILTON have presented a valuable addition to the knowledge of longitudinal vibrations of ship propulsion systems and their solution.

The surprise about this paper is that in these days a ship would be built whose natural frequency of shaft longitudinal vibration, uncomplicated by resonance in the machinery space bottom, would correspond to the blade frequency at full power. Even in open water operation, the longitudinal displacement of about 18 mils at the maximum operating speed would generally be considered'large. This is particularly true because of the high frequency of the vibration. It is a relatively simple matter in the early design stage to investigate the possible range of natural frequencies of longitudinal vibration of the propulsion system as a function of the range of stiffness values suggested in Reference 121 of the basic paper. In this case, the calculation of thrust bearing stiffness corresponds reasonably well with the range given by KANE and MCGOLDRICK. Because of the shallow floors under the thrust bearing, it could be anticipated that the foundation would be flexible.

As the design progresses, more definitive values of the thrust bearing stiffness and the foundation stiffness can be determined. In 1974, LITTLETON RESEARCH developed a computer program for Waukesha Bearing Corporation which performed all the calculations necessary to determine the thrust bearing stiffness, except the stiffness of the case.

Since the spacing of the bulkheads at Fr. 143 to Fr. 171 is less than 40 feet, the calculation of stiffness of the bottom on a static basis (i.e., neglecting the dynamic effects of machinery, structure and water interia) will probably be acceptable. A simple calculation of the type suggested in Reference [5] of the basic paper used in conjunction with MIL-STD-167-2 would have assured that the vibration experienced did not occur.

As suggested previously and so clearly shown in the paper, the difficulty with this installation arises because of the flexibility of the shallow double bottom. With the motor adjacent to the thrust bearing, this is difficult to remedy. A common solution, which frequently works very well, is to locate the feet for the thrust bearing on the lane of the centerline of the shaft and the foundation pads in holes cut in short longitudinal bulkheads on either side of the shaft. These longitudinal bulkheads extend from the inner bottom to the deck above and have the effect of restraining the rotation component which contributes so strongly to foundation flexibility.

In many installations, the stiffness of the thrust bearing becomes limiting. As pointed out by the Authors, this flexibility is governed by the buttons under the thrust shoes and the deflection in the levelling links. It might be of interest to note that the thrust bearing on the SL-7 ships have neither buttons under the thrust shoes nor levelling links. The thrust shoes have a radial cylindrical surface that bears against a plate inside the casing. The stiffness of the SL-7 bearings is on the order of 100 X lob Ibs/in.

For this ship, a static analysis of the stiffness of the thrust bearing foundation is adequate for the prediction of shaft longitudinal resonances. In large and high-powered Merchant


Ships, this may not be adequate. The resonant frequency of the second mode of vibration of the machinery space bottom, with the mode line across-ships near the middle of the machinery space, sometimes fall within the blade frequency range. Since this mode of bottom vibration couples well with the rotation of the thrust bearing foundations, the longitudinal shaft resonance can drop to the frequency of the bottom and the resulting vibration, being of wide extent, is particularly annoying. For predicting this type of trouble - and taking steps to avoid it in the design stage - quite sophisticated finite element analyses of the machinery space are required.

The Authors have presented a very interesting paper on the particular problems of longitudinal vibration in a class of Icebreakers and the steps being considered to correct the dif- ficulties. I appreciate the opportunity to add my comments and to commend the Authors for their work.

By: Mr. S. Karve. Dr. J.G. de Oliveira. and Dr. P.C. Xirouchakis

Department of Ocean Engineering Massachusetts Institute of Technology

We would like to congratulate the Authors on this interesting paper. The Authors state that typical open water centerline shaft axial stiffness values calculated using the recorded shaft thrust at the bearing and the measured bearing displacement fell between 5.5 X 10' and 5.6 X 10' Ibflin., depending upon the power level. For an elastic material the stiffness of the structure is determined by its geometrical and material properties and does not depend upon the level of excitation. It seems that the figures quoted above contain inertia effects and not static forces alone. We would like to ask the Authors: What is the reason behind the reported dependence of stiffness upon the power level?

Regarding the selected boundary conditions, it is a common practice in finite element modelization of ship structures to use first a coarse mesh that would include a substantial portion of the ship in order to determine more realistic values for the boundary conditions for a smaller finer mesh model. In the present case, the coarse model could include, for instance, the whole machinery space since this would give a better represen- tation of the whole structure. Would the Authors please comment upon this paint?

Regarding the computation of the flexibility of the thrust collar, we would like to ask the Authors this question. What were the boundary conditions used in their calculations and the main dimensions of the thrust collar?

Regarding the reported calculations of the bearing housing stiffness, we would appreciate it if the Authors would comment upon what they mean by discontinuity analysis used to compute the deformation of the forward end due to rocking and extension.

Regarding the methods for improving the vibratory behavior by changing the system stiffness, the Authors mention three areas for possible changes. The present paper only discusses alterations in the foundation design. Were other possibilities. involving in particular the shaft and the thrust bearing design, also considered?

In the FEM formulation the Authors state that they have used triangular and quadrilateral plate elements. The designation of plate elements needs some clarification. In fact, in the case of triangular elements the bending stiffness was made zero so that the element becomes a constant-strain triangular element for which no out-of-plane deformations are possible. Thus the designation of plate element should not be used. In the case of quadrilateral elements the Authors do not indicate whether bending stiffness was retained or not in the formulation. Would


the Authors please comment upon this point? If bending stiffness was not neglected, would you please describe how these elements were used in the model?

It will certainly be very useful to compare the numerical results obtained by the Authors with the experimental results obtained from a structural model, and the Discussers hope this comparison can be included in thefinal version of the basic paper.

By: Mr. Eugene C. Haciski Design Branch Naval Engineering Division U.S. Coast Guard Headquarters

This paper makes an excellent contribution to the solution of vibration problems of propulsion shafting systems for Icebreakers.

Regarding future Icebreakers, it would appear that the design philosophy for their propulsion systems should be based upon the concepts of “average service conditions” and “extreme service conditions.” From our experience with the POLAR Class Icebreakers it follows that propulsion systems for this type of ship should be designed for “extreme service conditions” such as that which occurs when the flow of ice in the stern area causes “ice processing” by the propellers. Simultaneously, every effort should be made to eliminate or at least reduce the amount of ice in the stern area.

In this connection, I also would like to suggest the following bold idea. Why not use specially reinforced. highly-skewed CRP propellers? By this means it may be possible to diminish ice impact and vibratory excitation forces and also to change considerably the excitation frequency.

As regards the stiffness of the supporting structure, including the foundation of the thrust bearing, it is evident that this must be consistent with the stiffnesses of the other components of the shafting system. To assure adequate stiffness for the hull structure configuration at the design stages, the thrust bearing loca- tion should not be based only upon the arrangements of the machinery and shafting system, but also on considerations involving the lines plan, adjacent hull structure, and machinery foundations.

The most expressive example of this problem is depicted in Figures B-2 and B-3 of APPENDIX B where the supporting structure of the thrust bearing foundations is limited below acceptable levels because of the physical extent of the hull shape. By the way, the Frame Number in these Figures should be 171 instead of 177.

By: Dr. S. Hylarides Netherlands Ship Model Basin Wageningen, The Netherlands

I would like to point out that this paper deals basically with one-half of the longitudinal vibration problem on the propulsion shafting systems of the POLAR Class Icebreakers and that this half is only partially treated. As the paper stands, one cannot but state that it forms an introductory study into the encountered vibration problem. I have to admit that this is also stated by the Authors. However, it takes some effort to find it. More importantly, it takes considerable experience in this field to find these statements and to understand them.

It must be appreciated that this vibration problem is fun- damentally a combination of hydrodynamic and mechanical phenomena. These phenomena have to balance each other, and for this reason a vibration problem is, in general, nor solved by mechanical measures only. Using a one-degree-of-freedom mass spring system, similar to that mentioned by the Authors

but subsequently discarded as being too simplistic, I will nevertheless try to explain my views and demonstrate that the hydrodynamics around the propeller can be of greater importance than the thrust block mechanics.

Using the sketch, shown herein as Figure 2, the response, “x,” to a periodic force, “F,” acting at frequency, “w,” is given by:

where, “oo,” is the system natural frequency. Assuming that the excitation has a constant amplitude, the plot of the response takes the familiar form shown in Figure 3(a).

In Figure 3(b), I have shown the plot for an excitation whose amplitude is frequency dependent and varying in such a way that the magnitude increases as the end of the frequency range of interest is approached. The resulting response is “xF(w)” which is the product of “x” in Figure 3(a) and “F(o)” in Figure 3(b). The result is shown in Figure 3(c). One may then be tempted to say that there exists a resonance at or just above maximum rpm. Considering the preceding remarks, it becomes clear that this may be an erroneous conclusion.

Objections to this line of reasoning could be voiced on the ground that the dependence of amplitude of excitation upon frequency may be exaggerated or that it is hypothetical. Al- though there may be some truth in this, this remark does not apply to propeller-induced excitation because the frequency dependence is in fact generated by propeller cavitation.

Nowadays it is well accepted that cavitation plays an important role by magnifying drastically the pressure fluctuations on the hull afterbody above the propeller. Similar effects can nevertheless occur with thrust fluctuations and the other dynamic propeller loads. Further, it .is general experience that controllable-pitch propellers and ice-reinforced propellers suffer more from cavitation than do conventional propellers. Thus, the magnification effects of cavitation on propeller-induced vibrations definitely must be considered.

Looking at Figure 3 of the basic paper we see that the increase in amplitude of the thrust block vibrations with rpm is rather abrupt, in fact steeper than could be expected because of resonance. It is then easy to imagine that close to full shaft rpm the propeller is cavitating suddenly over a large area and thereby affecting thrust fluctuations in a strong manner. It follows that the solution of this problem requires consideration of the effects of propeller cavitation.

Under icebreaking conditions the speed of the ship is reduced and the propeller is loaded more heavily. This in turn makes it more susceptible to cavitation and hence to stronger vibrations. 1 regret I cannot estimate the effect of ice in the propeller race.

f

I I Figure 2.

Naval Engineers Journal. October 1980 107


prepared in order to assess their relative stiffness in relation to the system axial behavior.

A lesson should be learned from these studies in any new Icebreaker design program to ensure adequate stiffness in the bottom structure in way of the propulsion units and thrust blocks.

Figure 3.

I would have preferred that the basic paper had started by stating explicitly such considerations, then eliminating one after another all possibilities, and then finally arriving at the conclusion that resonance is the most probable cause.

Nonetheless, when focusing our attention on longitudinal stiffness, I would like to compliment the Authors on their com- prehensive report on the use of the finite element technique. Their exposition clearly shows that the finite element method has matured and it finally can be accepted as a marvellous analytical tool. For this contribution, I would like to thank them very much.

By: Mr. C. Klop and Mr. S. Basu German & Milne. Inc. Montreal. Canada

In this paper the Authors use a finite element analysis to investigate longitudinal vibration problems on the USCG Icebreaker Polar Stor in the shafting systems thrust bearing foundations and adjacent structures.

Three finite element models were used in the NASTRAN Computer Program, and the results of this investigation can be used as a tool in understanding Icebreaker vibration problems.

This analysishad onegreat advantage since an actualship was used and all st~uctural drawings were available to be used in the program. This analysis was made to determine not the axial stiffness of the shafting system itself, but the stiffness of the structure which transmits the propeller thrust into the ship.

This method could be used successfully in new designs provided fairly detailed structural stern arrangements were


REPLY BY THE AUTHORS

1) To Mr. Norman 0. Hammer-

The discussion submitted by MR. HAMMER is very much appreciated and in response to his opening remarks it must be stated that following the initial vibration reports the “behind the scenes activities” were really not as entertaining or interesting as one might envision. Most of the discussions were technically oriented and the objectives in all meetings have been the understanding and solution of the engineering problems.

The question of “whether or not there is a problem” was raised early in this project. Today we have reached the firm conclusion that there exists a serious problem when the propellers are processing ice. Although the open water data in our Figure 3 may not seem alarming, recently analyzed data from full-scale trials in ice have shown that the thrust bearing movements have amplitudes of the order of f 100 mils. Despite the fact that the mission profile may not always include a high percentage of high-power or ice-processing operations, those instances when these situations exist are extremely harsh and wearing on the machinery plant. For example, about fifteen percent of the casualty maintenance costs expended to date can be attributed to excessive vibrations, and this fraction is forecasted to grow as more expensive casualties (in the main motor gear) are experienced. Until such time as failure analysis permits the definition of suitable performance indices, the question of what is acceptable vibration under ice operations will not have been answered.

The “textbook” type of approach criticized by M R . HAMMER is used in our paper only for the purpose to identify systematically the tasks to be performed in correcting the problem. It is still believed that one of the first things to do is to alter the propeller excitation, even though water solutions are known to be ineffective when milling ice or receiving head-on blows from chunks of ice. It may be possible to achieve some improvement with certain geometric changes, but the best approach probably will be to ease the inflow or to avoid the chunks as much as possible. It still appears to be very difficult to alter the frequencies (or frequency bands) over which the excitation seems to be acting, and the Authors were the first to recognize that altering blade number or shaft speed are not useful solutions because the problem is not similar to those due to propeller excitation in open water. To add to the above, the cost factor is a major determinant of what should be done, and the risk of failure is ever present, as M R . HAMMER points out so well.

In replying to MR. HAMMER’S assessment of our recommendations, the Authors would like to point out the following:

I ) The use of full-scale measurements is an integral part of the overall approach. Both theory and measurement must be used in this case and the efforts along both lines are continuing.

2) The recommended analytical studies have as their objective the understanding of how the bearing housing behaves and how it can be altered to increase substantially the bearing stiffness; to understand the factors involved in yielding natural frequencies that cannot be predicted on the basis of linear theory; to explore the influence of alternate forms of dynamic ab- sorbers; to investigate the effects of possible coupling problems

SESSION NO. 6-PAPER #1 ASNE DAY COMMENTS

between torsion and extension; and to have tools that will permit systematic parametric studies where the effects of changes can be tested before steel is ordered.

3) We feel that changes may have to be made to all parts of the system as well as the excitation in an attempt to reduce the levels of this type of vibration. Time limitations and the lack of suitable data did not permit us to deal with anything more than one aspect of the overall problem in our paper.

2) To Cdr. William M. Simpson, Jr., USCG-

The Authors would like to thank CDR. SIMPSON for his kind remarks and for his interesting contribution on oil film stiffness.

The film flexibility quoted in our paper corresponds to steady state operation at specified thrust and speed conditions. During icebreaking, both the thrust and speed change with the passage of time, and the apparent flexibility of the film will alter in a manner dependent upon the actual thrust and speed time histories. The calculations submitted by CDR. SIMPSON appear to yield a plausible stiffness reduction for the total bearing stiffness for the assumed conditions used in his example. It is not possible, however, to make any comment upon the accuracy of his prediction because he has neglected the inertial (or dynamic) effects associated with these changes.

It is the Authors' view that the eventual solution of longitudinal shaft vibration during icebreaking will have to be based upon nonlinear models of the system which employ the exact force-deformation relations, not only for the oil film but also those bearing internals that behave in a nonlinear fashion. This problem is currently under investigation, and it is hoped to have some results available for publication in the not too distant future.

3) To Mr. John R. Kane-

The Authors were especially pleased to receive comments from such a distinguished authority as MR. KANE who was in- timately involved with similar problems on U.S. NAVY Bat- tleships in the past and who was a co-author of one of the classical papers on the subject of longitudinal shaft vibrations.

The points raised by MR. KANE on the extraordinary stiffness contributions of shaft, bearing, and foundations for these ships, the "sticktion" effects in disconnect couplings, the random loadings due to ice-breaking, and the multitude of critical speeds present here, all serve to further highlight the complex- ities associated with the Icebreaker vibration problems.

The possibility of axial-torsional coupling was raised very early in this investigation. At present the influence of this mechanism is being investigated on the basis of accurate predictions of the coupling parameters due to the propeller.

As regards the question of how the thrust bearing was modeled, we would like to assure MR. KANE that the nonlinear contact deformation was properly taken into account for all elements within the bearing that generate "stiffness" by such mechanisms. We wo not believe, however, that it is necessary to make arbitrary decisions on how the slope should be defined since nonlinear techniques can handle this problem exactly.

Finally, it is comforting to hear from MR. KANE that on the basis of his experience, our predictions and conclusions regarding foundation stiffness are "probably quite correct."

4) To Capt. L.C. Melberg, Jr., USCG (Ret.)-

The comments provided by CAPT. MELBERG, who was involved not only in the early design stages but also with the maintenance of these Icebreakers immediately following con-

struction, are especially informative and serve to place our paper into perspective. His confirmation of the seriousness of the vibration problem and its present and future effects on the machinery plant are appreciated, and it puts to rest the question of whether or not some improvement is necessary. In response to his question regarding the owner's intentions, we refer him to the short-term and long-term objectives that have been established on this subject and reiterate that the COAST GUARD is committed to making improvements.

5) TO LL Robert W. Gdick, USCG-

The Authors wish to thank LT. GULICK for his kind remarks on the value of the paper. We also welcome his comments that summarize the approach followed in the past by the U.S. COAST GUARD in resolving this problem and that outline the current and future courses of action that have been established.

6) TO Mc. T.H. DOIISS~II-

At the time our paper was written the full-scale data obtained by DTNSRDC had not been analyzed, and so it was not possible to include a comparison of theory and experiment. It is hoped that the good agreement mentioned earlier between the computed stiffnesses in our paper and those supplied by several Discussers from preliminary data based upon the DTNSRDC finite element studies, model tests, and full-scale trials will provide the validity that MR. DOUSSAN feels would be necessary to make this study credible. The data and guidelines submitted by MR. DOUSSAN on the basis of his experience are very much appreciated.

7) To Dr. R. Weddsma-

The discussion provided by PROFESSOR WERELDSMA has been received with great interest because of his cogent treatment of the major aspects of the shaft vibration problem.

His assessment that the resonance phenomenon may not be the essence of the problem coincides with our opinion. Likewise we emphatically agree that the stiffness of the subject foundations cannot be increased significantly without making drastic structural and configuration modifications. As mentioned earlier, we have begun a careful study of the axial-torsional coupling problem because there are indications that it may be important, especially in the course of icebreaking when clearances in the thrust bearing, disconnect coupling, and gear- box alter the system properties.

For quick or preliminary estimates of the overall axial stiffness of the shaft, thrust bearing, and foundations, it is entirely permissible to neglect the contribution of the oil film. This is indeed in contradistinction to the case of coupled transverse shaft vibrations where the associated bearing coefficients must be taken into account as accurately as possible.

8) To Monsieur Guy C. Volcey-

The Authors are extremely pleased by the favorable response that our paper received when it was reviewed by a well-known authority such as MONSIEUR VOLCY. It is indeed gratifying to note his agreement on the philosophy and methods used to at- tack the problem, the decisions taken in the modeling process, and the design alterations proposed to correct the problem.

In response to his questions we would like to inform him as follows:

a) The "y" and "z" axes of our Figure 4 were inadvertently ntislabelled when the sketch was prepared for the basic paper.

Naval Engineers Journal, October 1980 109

ASNE DAY COMMENTS SESSION NO. 6-PAPER #1

When these are interchanged it will be seen that the boundary conditions stated in the paper are correct.

b) The thrust bearings on these ships were manufactured by WAUKESHA and have their flanges at the shaft centerline level.

c) Within linear theory there is no difference between tilted and untilted conditions, and so all calculations were made for the untilted condition.

d) The axial damping of the thrust bearing was predicted using lubrication theory. Foundation damping has not yet been investigated.

e) The use of struts was considered in the case of the “reinforced” foundations but it was rejected after it was found that they did not offer any substantial increase in stiffness because of the flexibility of the shell plating.

The Authors are most appreciative of the valuable data and suggestions that MONSIEUR VOLCY has submitted in his contribution, and we look forward to the opportunity of addressing in the near future the many questions and unresolved problems that he mentioned.

9) To Mr. F. Everett Reed-

The Authors wish to thank MR. REED for his kind comments and valuable suggestions.

The question of whether or not a simple static stiffness suf- fices in this case has not been answered conclusively, especially since the ship is surrounded by ice, and for this reason, the Authors would like to explore further the inertial effects involved when the propeller vibratory energy is diffused within the ship. In this respect, one should, without hesitation, place these Icebreakers in the class of large, highly-powered Merchant Ships where sophisticated finite element analyses of the im- pedance of the machinery spaces are required.

It is particularly reassuring to have MR. REED‘S comments on the SL-7 bearings because it is strongly felt that the Icebreaker bearings can be re-designed to have, among other things, an axial stiffness which is approximately ten times larger than the current value.

In retrospect, it is, of course, always easy to criticize the designers of a ship when problems arise in service. In the case of these Icebreakers it is probably fair to say that the state-of-the- art was not as developed as it is a decade later, and that the magnitude of the ice forces had never been clearly determined so as to permit rational design decisions.

10) To Mr. S. Karve, Dr. J.C. De Oliveira, and Dr. P.C. Xirouchakis-

In response to the questions raised by MR. KARVE and DRS. DE OLIVEIRA and XIROUCHAKIS, the Authors would like to re- spond as follows:

a) Aside from the fact that the measured “stiffness” may be contaminated by inertial effects, the apparent axial stiffness of a thrust bearing varies with shaft speed and transmitted thrust because there are several nonlinear elements within the system.

b) The question of boundary conditions was dealt with in our response to MR. ANTONIDES on page 81 of the August 1980 Journal.

c) The thrust collar was fixed at the shaft. The collar outer and inner diameters are 27.0 in. and 13.0 in. respectively, and its thickness is 9.5 in.

d) For a detailed treatment of the subject of “discontinuity” analysis, the Discussers are referred to Chapter 15 of Reference [ 1 ) cited herein.

e) For the existing Icebreakers, shaft alterations are extremely difficult and costly and were therefore discarded. In contrast, the Authors are vigorously pursuing the possibility of re-designing the thrust bearing.

0 For a detailed treatment of the properties of the CQUAD4 and CTRMEM finite elements used in this study, reference is made to the NASTRAN Theoretical Manual 121.

g) Preliminary correlations between theory and measurement have been mentioned in earlier contributions by other Discussers.

REFERENCES

[ l ] Timoshenko, S. and S. Woinowsky-Krieger, Theory of Plates and Shells. New York, N.Y.: McGraw-Hill Book Company, 1959.

121 MacNeal. R.H. The NASTRAN Theoretical Manual. The MacNeal-Schwendler Corporation, December 1972.

11) To Mr. Eugene C. Haciski-

The Authors welcome and concur with the comments provided by MR. HACISKI and are pleased that now they will be available for consideration by future Icebreaker designers.

As regards to the use of highly-skewed propellers, it can be said that ongoing and future propeller/ice research here and abroad may possibly validate MR. HACISKI’S ideas. However, as in the response to MR. DASHNAW, page 81, August 1980Jour- nal, it is felt that “open water” solutions may not be applicable to ice processing.

The Authors wish to thank MR. HACISKI for pointing out that in our Figures 8-2 and B-3 of the APPENDIX the frame number must be 171 instead of 177.

12) To Dr. S Hylarides-

DR. HYLARIDES has underscored the limited scope of our paper but no apologies are offered, for reasons cited in earlier replies.

The Authors believe that the explanation submitted by DR. HYLARIDES on the subject of frequency-dependent propeller loads caused by cavitation is very helpful for ship operation in open water. Nevertheless, all of the measured data on these Icebreakers have indicated that vibration is not a problem except during ice-processing. Under these circumstances it is difficult to see how cavitation can play a significant role when blades mill ice or when they suffer severe impacts from ice.

The Authors are aware of the early use of finite element techniques by DR. HYLARIDES and are therefore pleased with his concurrence that the technique finally has matured for routine design work.

13) To Messrs. C. Klop and S. Basu-

The Authors agree wholeheartedly with the summary provided by MR. KLOP and MR. BASU in that any new Icebreaker design program should include a detailed finite element analysis of the structure in way of the thrust bearings and other propulsion system components. In fact, one may go further and state that the available technology suggests that any major new ship design should be examined in similar fashion. In this respect, it is hoped that our paper has contributed toward the development of this new trend.


session no. 6 — paper #1: comments and discussion — part ii

Documents