Fatigue of ship structures

STUDIECENTRUM T.N.O. VOOR SCHEEPSBOUW EN NAVIGATIE Netherlands' Research Centre T.N.O. for Shipbuilding and Navigation

SHIPBUILDING DEPARTMENT MEKELWEG 2, DELFT

FATIGUE OF SHIP STRUCTURES

VERMOEIING VAN SCHEEPSCONSTRUCTIES

by

IR. J. J. W. NIBBERING

(Senior scientific officer of the Ship Structures Laboratory Delft)

Issued by the Council

This report is not to be published

The influence of variations of loading condition . . . . Local bending of longitudinal

temperature and changes in 10 members, of the bottom due

,

..

. 12

Actual peak to peak values; fatigue loading line

9 The fatigue 'strength of the "Canada'..

Conclusion and final observations, References,

page

Summary .

,

3

1 Introduction.

,2 'General opinions,

practical. experience and results of

fatigue tests .

.3 Dynamic loading of the longitudinal structure of ships 5

4 Slamming . .

5 Corrosion allowances and tolerances of thickness

,

10'

to waterpressure . . .

...

. n 1, II 3 3 7 . . 6 7 8 12 15 10 18 20

FATIGUE OF SHIP STRUCTURES*)

by

IR. J. J. W. NIBBERING

Summary

Fatigue is often considered to be a problem of secondary importance in naval architecture. The correctness of this opinion will be investigated in this article. For this purpose the loading of the longitudinal structure of a modern dry

cargo ship is analyzed.

Attention is given to slamming, temperature stresses and local stresses due to waterpressure. In this article any

overestimation of the fatigue loading has been avoided. As a consequence several procedures and calculations could be

kept simple. The resulting "minimum" fatigue loading of the ship proves to be rather severe.

Some general conclusions about structural design are presented.

1 Introduction

For a long time naval architects have considered

fatigue to be a problem of minor importance in

shipbuilding. Often this conception will not have been based on rational knowledge but merely on the feeling that where ships are moving slowly in

comparison to aeroplanes, trains and cars, the

number of load cycles during a ship's life cannot be very high.

The disregard of fatigue can also be caused by the fact that research in the field of brittle fracture was of primary importance for many years. Anoth-er possibility is that practical expAnoth-erience has not given indications for the desirability of the study of fatigue phenomena in ships.

In this paper an attempt is made to find the

proper place for fatigue problems in ship research.

2 General opinions, practical experience and

results of fatigue tests

In the design of steel structures such as bridges and cranes, the fatigue aspect of loading is taken quite seriously. Often a fatigue life of at least 2 million cycles is required. This practice differs greatly from what is usual in shipbuilding where calculations for longitudinal strength are purely statical in

char-acter (ship in standard wave).

A comparison between the loading of a railway bridge and a ship can give a first impression about

the practical value of both points of view. A bridge with 40 trains passing each day will endure about 2 millions of load cycles in 100 years. The number of changes of the longitudinal bending moment in a cargo ship in very rough sea (> 7 Beaufort) will amount to 10.8 millions after 30 years [1]. But it

must be mentioned that the corresponding

lon-gitudinal stresses belonging to it are surprisingly low as can be seen in figure 1.

This diagram has been presented by Dr. YUILLE

[2]. The straight lines for "one year" and "30

years" represent the cumulative frequency distri-bution of the longitudinal bending stresses for the well known "Ocean Vulcan".

If we look at the line "30 years" in figure 1 we

find that only 10, of the above-mentioned 10.8 million load cycles have double amplitudes of

stress 1) higher than 400 kg/cm2. Only 100 cycles of stresses surpass 800 kg/cm2. In figure 1 Dr. YUILLE

has given a special curve for the stresses at

dis-continuities with a stress concentration factor equal to 3. This curve nearly touches the fatigue-strength curve for notched specimens in corrosive

environ-ment. This is in conformity with the fact that

during the tests with the "Ocean Vulcan" new

cracks did not appear and existing small cracks did not become larger.

In France JOURDAIN [3] has come to a similar conclusion. In practice many fatigue cracks have been found. VEDELER for instance does not doubt

*) Publication no. 12 of the Ship Structures Laboratory of the Technological University, Delft For convenience, this value will be called "stress" hereafter.

,

TYPICAL FATIGUE LIFE

I NOTCHED SPECIMENS UNDER CORROSIVE

I

that most of the cracks found in ships are due to fatigue [4]. STENEROTH [5] 1.s of the same opinion. Various relatively small cracks have been described by BOYD [6]. These were called brittle but nothing

proved that fatigue has not played a role in the

process of their formation.

Now and then the International Institute of

Welding-Committee XIII delivers reports in which fatigue cracks in actual structures are discussed.

Report I.I.W./14-59 [7] deals with cracks in the

forepart of a ship immediately above and parallel

to the connection of the horizontal tanktop and

the side shell plating. These cracks were no doubt due to slamming.

In a tanker 129 cracks have been found; a large

portion was situated in bulkhead stringers. The

tanker had been at sea for 41/2 years. In figure 2

FATIGUE LIFE

PLAIN MS SPECIMENS

\

FATIGUE LIFE

PLATES WITH BUTT WELDS

CONTAINING FLAWS

CONDITIONS

(left) a typical crack can be seen. At the end of the report the committee presents the following con-clusion:

"Straining the corners of the tank so as to open the angle can be explained by the effects of inertia of

the liquid cargo in the tanks, rather than by the

general bending of the whole structure in a swell.

The welded joints between the brackets and the flanged edges of the stringers seem to present a

markedly reduced resistance to this kind of stress,

perhaps partly because the bracket is not in the

same plane as the stringer and therefore there is a tendency, as the angle becomes more obtuse, for

bending of the bracket to occur around its point

of attachment."

Of course the main cause of the crack is the stress

concentration at its origin. But the fact that the

N.

0.35 70D3 -0.3 6000 -L.L1 0.25 Lu 5000 -0 u 0.2 \,-) 4000 - _ct UJ tri cr, cr Cl. --Z 3000 - LL 0.15 0 Lu cr UJ CC _J 2000 0 1000 - 0.05 10' 104 10S 106 o7 08 io9 10" NUMBER OF CYCLES

Figure 1 Diagram according to Dr. Yuille [2]

1C1 10

material of the stringer has been cold-bent and then welded to the knee will also have been

un-favourable.

In connection with the above-mentioned reports on cracks, two types of connections of bulkhead

stiffeners to stringers are shown in the right of figure 2, which have been tested in fatigue [4].

In practice the connection with a small knee had often shown cracks at the points indicated

caused by fatigue loading. The results of the tests

in which the stiffeners had been loaded in alter-nating bending are shown in the diagram. The

connection on the right proved to be much better than the connection with a small knee.

The foregoing discussion seems to indicate that the dynamic loading of bulkheads in tankers

de-serves much more attention in ship research. So far the problem has been treated-only in Japan

[9] [10] and Norway [8]. Fatigue testing of ship structural components has also obtained relatively little attention for many years. WECK [21] inves-tigated the bending fatigue strength of various

con-nections of stiffeners as well as the strength of scalloped stiffeners. NEUMANN, JACKWITZ, MaLLER

and others worked in the same field and also tested bilge frame connections and connections of contin-uous longitudinal frames to transverse webs and bulkheads [22] [23]. De. LEIRIS and DUTILLEUL investigated interconnections of longitudinal frames at tranverse bulkheads [20]. Finally JAEGER and NIBBERING have tested various connections between orthogonally placed stiffeners like beamknees [18].

In the latter investigation many specimens were

Figure 2

loaded in such a way that cracks developed after a relatively low number of cycles. This low cycle fatigue was thought to be more representative for ships than ordinary high cycle fatigue.

A discussion of the results will not be given here but figure 3 gives a general impression. From the figure it can be concluded that cracks as indicated

in the left of figure 2 can be avoided by the

ap-plication of well rounded knees like no. 3 and 4.

3 The dynamic loading of the longitudinal

structure of ships

In connection with figure 1 it has already been

FATIGUE BENDING TESTS [4]

I bk .1101101 MalinUMMEMEMINEEI

minompon

IIlIIIIIIIIIIIIHhIl

3 NUMBER OF CYCLES

said that the dynamic loading of a ship in seaway

does not seem to be very severe. The diagram

however has not the pretension to be very accurate. The proof of Dr. YUILLE'S opinion that the danger

of fatigue in ships is small, is more particularly provided by the favourable experience with the

"Ocean Vulcan".

As a consequence this opinion is only valid for

ships of that kind and possibly not for modern

cargo ships. That is why in this article use is made of R. BENNET'S extensive measurements on two fast cargo ships: the "Canada" and the "Minnesota",

which are of the closed shelterdeck type. The dead-weight of of the "Canada" was 9,085 tons; the service speed was 19.5 knots. The ship sailed between the

Channel and the West Indies. The "Minnesota"

sailed in the North Atlantic region; her deadweight was 7,260 tons and her speed 19 knots.

The cumulative frequency distributions of the

SMALL CRACK

17

10

45

DOUBLE AMPLITUDE OF STRESS (G ) KG/rnm2 113 ,2x103 nr 9x103 n.2 5 x104 n.6.5x104 1 n.8.4 x106 1x103 1x104 1x105 Figure 3

N NUMBER OF TIMES THAT THE INDICATED STRESSES HAVE BEEN EXCEEDED.

Figure 4

lx 106

27A

THE POSITION OF THE STRUCTURES (HORIZONTALLY, REFERS ON, TO THE FATIGUE LIFE OF THE DETAILS INDICATED BY ARROWS

25 20 15 10 5 0 108 N r1.1 x107 _ HORIZONTALLY WELDED U NMACHINED _ _ __ ,... CA4/1; _ 41* All 6, 7,tf , 1:4 7 .. _ '..4'..'....Z., 1 _ I T 10 102 103 104 105 106 107 25 24 23 22 21 20 19 18 17 16 15 14 13 12 11 10 9 8 7 6 5 4 3 2 12 1x1071

-longitudinal bending stresses can be seen in figure 4. In this diagram BENNET has drawn a straight line

parallel to the lines obtained experimentally in

order to arrive at the frequency distribution for a lifetime of 20 years.

Comparing figure 1 with figure 4 it is

remark-able that the loadspectrum of the "Canada" for

1 month seems to be as severe as the loadspectrum

of the "Ocean Vulcan" for 30 years. This must

be traced to the differences between both ships in speed (7 knots), dimensions, form and type. The blockcoefficient of the "Ocean Vulcan" was 0.75;

the "Canada" had 0.65. The "Ocean Vulcan" is a full scantling ship. The "Minnesota" and Can-ada" are of the closed shelterdeck type with

ex-tended forecastle, which permits a relatively large amount of cargo to be loaded at the ends. Due to this, large inertia forces can occur during pitching and heaving which will influence the longitudinal

bending stresses. It should be observed that this

influence cannot be taken into account in a static standard wave calculation.

Although the stresses in the Swedish ships are relatively high, there still seems to exist little danger of fatigue as can be seen in figure 4. In this figure a typical general fatigue-strength curve for a hor-izontally-welded V-joint is included (see also sec-tion 9). Comparing this curve with the loading of

the "Canada" it appears that if the highest stress

(15.2 kg/mm2) in the "Canada" would occur two

million times instead of once, this V-joint still would not crack. Actually the situation is not so simple. This has already been demonstrated in

figure 1 where stress concentrations and corrosion fatigue have been taken into account. Before dis-cussing their influences it is first necessary to apply

corrections to the nominal longitudinal bending

stresses. These corrections are related to: Slamming and whipping;

Changes in temperature and loading conditions;

Changes in waterpressure on the bottom in

seaway;

Deviations in plate thickness due to corrosion and steelwork-tolerances.

4 Slamming

Slamming has in common with fatigue that widely

divergent views are enunciated about the role it

should play in strength calculations for ships. It is generally agreed that slamming stresses should be taken into account when brittle fracture or collapse of a structure is discussed (extreme load design). The high frequency of the slamming stresses,

how-ever, is thought to constitute an extenuating factor. This is disputable, for a high frequency can be an

unfavourable factor as well, if brittle fracture is

considered.

Besides, the frequency of the vertical two-node vibration is not always so very high; in long and slender ships a period of one second is quite possible. The conception that slamming should be included in the load spectra of ships is often opposed with

the argument that each captain will avoid

slam-ming by reducing speed and/or changing course.

This is only true if we say "try to avoid" instead

of "avoid" and "severe slamming" instead of

slamming.

In the light load condition it is often impossible

to avoid slamming; to a lesser extent this also

applies to the semi loaded condition. Now if

slam-ming in the light load conditions should be

tol-erated there should be no objection against a cer-tain amount of slamming in any other condition.

Slamming is taken here in a broader sense than

is usual. Any load causing vertical two node vibra-tions will be called slamming for convenience. The

well known "bangs" on the fore and aft bottom

parts will be called slamming as well as vibrations due to bow flare immergence.

It may be useful to quote here a few sayings of investigators who are familiar with the phenomena of slamming and whipping.

a. Korvin-Kroukowski ([13] pages V36, V40, V41) 10 Slamming of slow ships:

". . . ships with full lines, bread bottom area at

the bow and small dead rise:

These ships are known to slam frequently in head seas when in light draft condition. The Ocean Vulcan slammed during one out of 3

days in open ocean under light load condition. The impact of the water on the ship's bottom is the most conspicuous part of the slamming. In a cargo-ship the prior emergence of the bow is a necessary prerequisite to slamming." 2° Slamming of fast ships:

"The submergence of the bow and the dynamic effect of the bow flare appear to be the major causes of slams in ships of a destroyer type." ". . .The bottom impact plays only a second-ary part and the bow emersion is not a

nec-essary prerequisite to the occurrence of slam. . ." ". .In addition to the severe shocks recorded

as slams, the frequent bow immersion in the

head seas causes shocks of sufficient magnitude

to maintain the hull in a continuous state of

vibration."

.

. .

. .

R. Bennet ([12] pages 13, 8, 9) :

Minnesota in very heavy seas running at 13

knots on the port bow.

. .

and later on the

starboard bow:

"The big peaks were certainly caused by power-ful bottom impact loads obliquely on the star-board bow."

"When speed was reduced to 6 knots there was

rather an increase in the great majority of

amplitudes as shown by the larger r-value, but there were no further extremes of the same size."

. No account is taken of the biggest

slam-ming peaks, but only of the normally occurring slamming."

"The maximum value includes frequently oc-curring slamming increases but not the

occa-sional, very big peaks which were observed

particularly on the Minnesota." G. Aertssen ([14] pag. 7)

Dry cargo-ship; closed shelterdeck, Dw =

11,000 tons; service speed 16 knots.

"To these longitudinal bending stresses are

superposed vibration stresses exited by hydro-dynamic pressures. These vibration stresses. . .

did not induce any reaction among ships's staff in the full load and nearly full load condition."

(The vibration stresses at their maximum are not in excess of about 20% of corresponding maximum vertical bending stress.)

"The duration of the vibration is 10-30 seconds.

There are vibrations every 5 minutes in the light-load condition and one in half an hour in the nearly full-load condition. At more or

less regular intervals, about every 7 minutes, heavy oscillations are superposed on the stress curves and these oscillations of a stress range

of 0.8 to 2 tonsisq.in when they start are

quickly damped to half their value after 6 to

6.5 sec, and they are extinguished after about

25 sec. They rise with a shock and ships's

officers often respond to them as they respond to uninterrupted racing of the engine by

re-ducing speed. It is shown by the records that

this phenomenon slamming is connected with bow emersion as it always is preceded by

one, two or three pitch oscillations of large amplitude, the last one being the strongest.

This slamming did not occur in full-load

condi-tion even when the ship with her full power

and a speed of 13 knots headed into waves up to H1 /10 = 18 ft. In light-load condition, how-ever, even in a sea about 11 ft high, slams set in and speed had to be reduced."

Much has been written about slamming, but it

seems that it has never been tried to correct the

cumulative frequency distributions of the longitud-inal bending stresses for slamming. This is not so surprising for many factors are involved varying from the mentality of merchant marine officers to

the number of voyages made in partly loaded

conditions. This really should constitute an addi-tional argument for collecting information in such

a way that low frequency stresses and high

fre-quency stresses are recorded as well simultaneously as separately.

The author thinks that this is already done sev-eral times, but that the results have not (yet) been published.

The present article is not meant to cover a

thorough analysis of slamming in order to be able

to correct the cumulative frequency distribution

of the longitudinal bending stresses very accurately. It only tries to make a justified estimation.

Recently a report has appeared which has been of great help in this respect [14]. It refers to

meas-urements carried out on the "Lukuga" ; a closed

shelterdeck ship of 11,000 tons deadweight. The service speed was 16 knots. From the quotations above it appears that in [14] a clear distinction is made between "vibrations" and "slamming" ("vi-brations excited by hydrodynamic pressures").

For the light load condition there are some data

about the frequency distribution of "slams" and

"vibrations". During a 40 minutes-test the highest slam-stress amounted to 1.8 kg/mm2. There were about 19 slams for 350 wave bending cycles.

More-over 21 vibrations occurred with a maximum of 0.6 kg/mm2. 40% of all slams were higher than

0.6 kg/mm2; 60% of all vibrations exceeded 0.2

kg/mm2 (see [14] figure 9). The wave bending stresses generally corresponded more with the

"Ocean Vulcan" stresses than with the "Canada" stresses, so this information cannot be applied to the "Canada" without some modifications because:

I. The "Lukuga" is a somewhat more rigid ship

than the "Canada"; as a consequence the gen-eral level of the "vibration" and "slam" stresses will be lower.

2. The speed of the "Canada" is 19.5 knots; the

speed of the "Lukuga" 16 knots.

In the light load and half loaded condition the largest slamstress of the "Canada" and

"Minne-sota" amounted to 50% of the highest wave bend-ing stresses durbend-ing a short test. In the fully loaded

condition this percentage was about 40%. The 73 tests of the "Canada" have been summarized

on page IV-12 in [1].

It can be deduced that:

The ship was fully-loaded in 25 tests The ship was light-loaded in 12 tests The ship was half-loaded in 21 tests

The ship was nearly fully-loaded in 13 tests. For our purpose it is allowable to assume that the

ship had been fully-loaded in 40 (25+15) cases

and light-loaded in 33 (12+21) cases.

In 31 of the 73 cases the maximum wave bend-ing .stress durbend-ing a particular test was found to

be smaller than the maximum of the sum of

wave bending stress and Slam stress. (Of course this does not mean that slamming has not oc-curred during the other tests.)

The largest slam stress exceeded 2.8 kg/mm2 and occurred in the half-loaded condition. (The highest slam stress. of the "Ocean Vulcan" was 2.25 kg/mm2.)

More information about slamming can be found

in [15] and [16] which refer to tests with 3 destroy-ers of the Royal Netherlands Navy. As can be seen

in table I derived from [16] 155 slams were re-corded during 2,752 cycles of the longitudinal bending moment. The table demonstrates that at

relatively low speeds (17 knots and 12 knots) slam-ming occurred in head, bow, beam and quartering

seas. A large part of the slams was not due to

"bottom impact" but to "bow flare immersion'. The ship with the smallest number of slams

(ship S) met 38 slams during 869 cycles. Most of these slams occurred in bow seas. It is reasonable to suppose that this situation will always be avoided by changing course so that bow seas will be changed

in head seas. For the 97 cycles of encounter for

which ship S sailed in bow seas at 12 knots another

97 cycles will be met in head seas. In that case,

according to Table I, only 3 slams would have oc-curred in stead of 7. Along the same line it can be

TABLE I Number of slams per run per ship in rough or X-sea state

deduced that at 17 knots only .3 slams would have occurred in stead of 16. On the whole 38-17 = 21 slams will be met during 869 cycles which is about 1 slam to 40 cycles. This number is thought to be representative for the "Canada" too for the follow-, ing reasons:

J. The speed of the "Canada" is as, high as 19.5 knots

.2. _{The small blockcoefficient and the presence of} extended forecastle for cargo probably resulted

in a certain amount of bow flare,

so, the "Canada" will be more or less comparable,

with the destroyers.

Generally the destroyers were fully-loaded while

during 45% of the time the "Canada" sailed in a partly loaded condition which is worse. In the partly-loaded condition the "Lukuga"

met 1 slam during 18 cycles.

With the assumption of 1 slam to 40 cycles for the "Canada" the influence of slamming will certainly not be overestimated.

If we take all the tests of the "Canada" together,,

the ship has experienced about 180,000 cycles..

From these, 180,000 : 40 = 4,500 cycles will have been accompanied by slamming or vibrations.

Table II presents a presumed frequency distri-bution of slam stresses. It is probable that slam-,

ming never took place at wave bending stresses

smaller than 1 kg/mm that occurred 130,000 times (see summary of "Canada" tests in [1]). This means

that the 4,500 slams have to be distributed over

50,000 cycles instead of 180,000 cycles.

The columns ( 1 ) and (4) in Table II have been derived from the frequency distribution line of the "Canada" for 1 month in figure 4. In column (5) the number of slams of the intensities of column (2) have been estimated. These slam-stresses are given in proportion to the maximum slam-stress

that.prob-Speed Heading

Ship A Ship B Ship S,

Slams _encounterCycles of Slams _encounterCycles of '

Slams j , Cycles of encounter 12 knots Head 7 120 9 95 3 1 97 12 knots Bow 110 110 2 , 105 7 97 12 knots Beam 4 105 2 99 2 89 12 knots Quartering 6 93 3 87 '0 75, 12 knots Following r 0 60 1 47 0 62 17 knots Head II _{19 .102' 19 112} 3 WO 17 knots Bow 14 112 9, 112 16 1001 17 knots Beam . ,5 '1031 7 97 7 1001 1 17 knots 'Quartering 0 '94 0 96 0 87 17 knots Following .0 71 i) 70 I 0 62 Total 65 9701 52 913 38 869 :

TABLE II

ably has occurred (50% of the maximum wave

bending stress). In column (6) the number of slams in (5) have been summarized. The resulting cor-rection points for slamming have been indicated in figure 5.

5 Corrosion allowances and tolerances of

thickness

When the experiments with the "Canada" were

carried out the ship was relatively new and in good condition.

As a consequence there will not have been a reduc-tion of the scantlings due to corrosion. If we assume a corrosion allowance of minimum 10% then after

a 30 years' life the stresses in the "Canada" will be 10% higher than originally. On an average

this will be 5% in these 30 years.

Local deviations of the mean thickness of the

plates and sections due to fabrication and corrosion

can easily amount to 5%. A total correction of

10% for both phenomena has been applied in

figure 5.

6 The influence of variations of temperature

and changes in loading condition

JASPER has measured stresses on the tanker Esso

Asheville [17]. He found that the variation of

the "still water stress" caused by differences in temperature between day and night could be of

the same magnitude as the largest wave bending stresses.

The influence of these variations of stress on the frequency distribution of the longitudinal bending stresses is not very large. This can be demonstrated in the following way. Suppose the wave bending

moment to change about 10,000 times a day. In

30 days there will only be 30 variations of the still water stress due to differences in temperature

be-tween day and night. If the diurnal stress varia-tions would be 4 kg/mm2, then the point a = 4; N = 2,000 of the "Canada" line in figure 4 will

shift to N = 2,000+30 = 2,030. This is quite

neg-ligible. On the other hand the diurnal variations of stress increase the diurnal maximum value of

the longitudinal stresses as can be seen in figure 6.

(1) (2) (3) (4) (5) (6) (7) Wave bending stress in proportion to maximum value of wave bending stress = S Slamming stress in proportion to maximum slamming stress = 0.5 S Wave bending stress ( la) + slamming stress (2) Number of

cycles for _Estimated wave stresses _{number of}

larger than indicated in slams Total number of slams Indication in figure 5

(la) (lb) column ( la)

0.9S a s 0.9 x 0.5S 0.7 x 0.55 I .25S I I 1 I I 1 0.85 a 0.95 0.9x 0.5S 1.25S

-

1 2 0.7 x 0.5S 1.15S 3 5 0.5 x 0.5S I .05S I ,-_. 3 8 0.75 a 0.85 0.9x 0.5S 0.7 x 0.5S

-1.055 E:-. 4 19 3 0.5 x 0.5S 0.955 8 90 0.3 x 0.5S 0.85S ,.-7) = 6 96 0.9x 0.5S

-

,

-0.7 x 0.55 0.95S

c

19 38 4 0.5 x 0.5S 0.85S c, 0cz, 0 30 68 5 0.3 x 0.5S 0.75S Lri' c),c,

,

30 98 0.5S a 0.65 0.9 x 0.55 0.7 x 0.5S 0.5 x 0.5S 0.755 50 148 6 0.3 x 0.5S 0.655 70 218 7 0.1 x 0.5S 0.555 80 298 0.3S a 0.5S 0.9 x 0.5S

-0.7 x 0.5S

-0.5 x -0.5S 0.55S 52 350 8 0.3 x 0.5S 0.45S 500 850 9 0.1 x 0.5S 0.35S 1,000 1,850 10 0.1S a 0.3S <-- 0.1 x 0.5S 0.15S 2,650 4,500 II

-. ,

-

-0.65 a 0.75

-

--

-

--

-

-,

-

-

--

I

DOUBLE AMPLITUDE OF STRESS) iKG/mm2

102 103

f.1 =NUMBER OF TIMES THAT THE INDICATED, STRESSES

Figure,6

The temperature stress in one day is equal to A: the maximum. stress-variation indicated by the wave stress, recorder is equal to B. The actual

largest peak to peak value is equal to C.

5 10

HAVE BEEN EXCEEDED,

Figure 5 Fatigue loading of deck

In order to estimate the influence of these diurnal peak to peak values on the frequency 'distribution roughly, the maximum temperature stress in one

month is supposed to be 5 kg/mm2; the lowest

value 2 kg/mm2. The low value will generally occur during rather bad weather when the wave stresses are high. The higher temperature stresses will gen-erally be combined with low wave stresses. How

both types of 'stresses should he put together is

obscure for the maximum values will generally not coincide.

The high temperature stresses are very rare. A diurnal stress variation larger than 4 kg/mm2 will

1 1 J 0((se, 'Jo ( . 4 1 I 1 4'1C . 1 e & It. eo 4₀ HORIZONTALLY WELDED;! UNMACHINED. b.111 .41e' ii b.../Ps Silk C4 4 0 ('''7 cell. Ni . ..,.. '';1'1111111ftlihe.. 106 107 8 35 34 33 32 30 29 28 27 26 25 24 23 22 21 20 19 18 '-17 16 15 14 13 '11 10. 8

7-6 5 4 3 2 35 30 20 _ 15 10 31 12 9 0.5x104 10 25 5 N.

on the average certainly not occur more frequently

than three times a month. On these days the

maximum wave stresses will not be higher than

5 kg/rnm2. The maximum possible peak to peak

variation is equal to 4+5 = 9 kg/mm2, but as

there is little chance that both stresses .coincide more realistic value is 8 kg/mm2.

Now it can be seen in figure 4 that there are al-ready 20 variations of stresses higher than 8 kg/rnm2; another 3 cycles of this magnitude do not change the diagram visibly.

If we now consider the small temperature stresses

we know that they can be accompanied by high

wave stresses in bad weather. This will, however, seldom occur, for a stress of 7 kg/mm2 will not be exceeded more than 50 times during a whole month (see figure 4). Moreover, these stress variations will

be limited to a few stormy days. Even then it is

highly improbable for the highest wave stresses to

coincide with the extremes of the temperature

stresses because of their low number. Perhaps once

a month a combined peak to peak variation of 7 kg/mm2 wave stress and 2 kg/mm2 temperature

stress will be possible. In figure 4 it can be seen

that again this has no influence on the frequency distribution. Probably this will also apply for com-binations of temperature stresses and wave bending stresses during the remainder of a month. In any case for the purpose of this article a more accurate analysis of the influence of temperature stresses is thought to be unnecessary.

It will be clear that the changes in the loading

condition of a ship will affect the frequency

dis-tribution of the longitudinal stresses to an even smaller extent than the temperature changes do,

because of the very low frequency.

In section 9 other aspects of this kind of stresses will be discussed.

7 Local bending of the longitudinal

mem-bers of the bottom due to waterpressure

and cargo weight

The cumulative frequency distribution of the

"Canada" in figure 4 applies to the stresses in the deck.

The structural material of the "Canada" is dis-posed in such a way that the wave stresses in the bottom are 75% of the wave stresses in the deck. Figure 7 shows the frequency distribution of the bottom stresses. The stresses introduced by

varia-tions in waterpressure and cargo pressure have

been included.

These loads can be roughly divided into :

The load caused by differences in hydrostatic pressure when the ship moves through waves

(figure 8: H m waterpressure).

Differences in hydrostatic pressure due to heav-ing, pitching and rolling (figure 9: (1-1, H2) m

watcrpressure).

Variations in the pressure of the cargo on the bottom during heaving, pitching and rolling as a consequence of the inertia of the cargo.

Similar variations in waterpressure

(hydro-dynamic load).

These components are interrelated in a complex

way. Phase differences exist between them. Moreo-ver, each one is a function of the length of the ship. There is a lack of information in this field which

renders the estimation of the stresses concerned very difficult. For the present it is thought to be

sufficient if only the components a. and b. will be looked into.

It is supposed that the favourable influence of

the Smith effect and of phase differences (c. and d. will counteract to a certain extent) will support this opinion sufficiently.

The draught of the "Canada" was 7.5 m. In

the stilwater condition the local bending stress will be about 6 kg/mm2. If we suppose a wave height of 5 m for a weather condition of 7 to 8

Beaufort, the local stresses will vary to an amount

of 5/7.5 x 6 = 4 kg/mm2 for the load sub a.

This load will be in phase more or less with the wave bending moment amidships only.

The load b. will be caused by pitching for a large part. The hydrostatic stresses involved are very large at the ends of the ship. However, in these parts the wave bending stresses are small.

The sum of wave stresses and the stresses sub b.

will reach a maximum somewhere foreward or

aft 0.5L.

The total local stress sub a. and sub b. now will be taken very arbitrarily indeed 1.2 times the stress sub a. alone : 1.2 x 4 = 4.8 kg/mm2. A straight line distribution of these stresses has been assumed. They have simply been added to the wave bending

stresses for the probability that at certain cross sections of the ship these effects have the same

phase is considered rather high.

8 Actual peak to peak values; fatigue loading

line

The lines obtained up to now in figures 5 and 7 still underestimate the fatigue loading in a ship.

This can be seen in figure 10 in which an arbitrary

small part of the records of the wave bending

DOUBLE AMPLITUDE OF STRESS (0-) KG/rnm2 35 34 33 32 31 30 29 28 27 26 25 24 23 22 21 20 19 18 17 16 15 14 13 12 11 10 9 8 7 6 4 3 2

"

Figure 8 13 5 7 UN MAK [1] Figure 10 35 30 25 20 15 10 5

k

\ tcs,

"10,\ k,,,, ''S'e\

*

0 07.7 0,9 -,,,, 4,?s, N

t

\ir ve o,, e. c., V4- o rei b 411, le Nc ^*,,x If e, No. ... St,S` ItN .

'

\\\,' 0 4. ci 0_4,46 /907-7-0A7 .)

lt

\ ",.

NX.96,,,, NNN,,N . NN.N\"N '

,\,

,

10 102 103 105 107 108 N.

N =NUMBER OF TIMES THAT THE INDICATED STRESSES HAVE BEEN EXCEEDED.

Figure 7 Fatigue loading of bottom structure

5

1

0 10

stresses of the "Canada" and the U.S. Coast guard Cutter "Unimak" are shown.

The part of the curve from the "Canada" as

indicated by the numbers 0 to 7 comprises 7 half cycles of the following stresses:

0 to 1 3.30 kg/mm2 1 to 2 1.05 kg/mm2 2 to 3 1.20 kg/mm2 3 to 4 2.60 kg/mm2 4 to 5 2.45 kg/mm2 5 to 6 2.45 kg/mm2 6 to 7 3.00 kg/rnm,

If we now consider the peak to peak value between the lines 0 and 3 (3.45 kg/mm2) we see that this

value is larger than any one of the individual

stresses 1, 2 or 3.

This double amplitude of 3.45 kg/mm2 is real and should be incorporated in the frequency dis-tribution.

The peak to peak value between 0 and 7 is even larger and amounts to 3.74 kg/mm2 which is 13%

higher than appears in the "Canada" line by the number 3.3 kg/mm2. The larger the number of cycles the larger this percentage can be. On the

average roughly 1 out of 10 cycles should be taken

about 25% larger than has occurred up to now

in figures 5 and 7.

We have now the actual cumulative frequency

distribution of the nominal stresses in deck and

bottom of the "Canada". The remaining question is: will this ship withstand the corresponding loads without developing fatigue cracks. This is quite a

problem for very little information is available

about the strength of structures or structural details loaded in a similar way as deck and bottom of the

"Canada". If completely reliable data should be

required one would be obliged to apply the same sequence of the fluctuating and still water stresses as well as the same frequency of the various kinds of fluctuating stresses.

This seems impossible but it should be

remem-This can also be put as follows: There are There are There are 9 stresses with an 90 stresses with an 900 stresses with an

bered that testing of this kind happens daily in

practice with all ships sailing at sea. The classifica-tion societies even collect records of these tests for instance, reports on cracking. The major difficulty is that as many variables are involved as there are types of ships, types of structures, shipping routes etc.

Perhaps it will once become possible to handle this problem systematically on a large scale with full use of statistics and computers. At this moment we can only try to use the results of ordinary fatigue tests on welded structural elements and details.

First it is necessary to realize that each point

on one of the stress lines in figures 5 and 7 means

that a stress larger than the indicated value has

occurred the indicated number of times. So these lines can not be compared with results of ordinary fatigue tests where each specimen is subjected to

a dynamic load of constant amplitude and

con-stant mean value.

One possibility is the use of one rule or another

which "translates" the random load of a ship in

a fatigue load of constant magnitude for a certain number of cycles. Such a rule does not exist. The

well-known Miner's rule (Zn/N = 1) cannot be

used because it only applies to "ordinary" fatigue loading where the number of cycles to rupture is

larger than 100,000. The loading of a ship is a

combination of ordinary or high cycle fatigue and plastic or low cycle fatigue. Much research with programmed fatigue loading and much informa-tion from practice is needed before any acceptable rule or procedure can be established.

At this moment we can only try to avoid any

overestimation of the loading of the "Canada" and

replace the load spectrum by a few "ordinary"

fatigue loads each of which is at least equivalent to that load spectrum.

From figure 5 we see that there are 10,000,000 cycles of stresses larger than 1.8 kg/mm2. They can

be presented in the following way:

100,000 - 10,000 stresses having a magnitude between 11.7 and 8.1 kg/mm2 1,000,000 - 100,000 stresses having a magnitude between 8.1 and 8.4 kg/mm2 10,000,000 - 1,000,000 stresses having a magnitude between 4.8 and 1.8 kg/mm2 10 - 1 stresses having a magnitude between 25.1 and 21.7 kg/mm, 100 - 10 stresses having a magnitude between 21.7 and 18.3 kg/mm2 1,000 - 100 stresses having a magnitude between 18.3 and 14.9 kg/mm2 10,000- 1,000 stresses having a magnitude between 14.9 and 11.7 kg/mm2

CC average" magnitude of 1/2(25.1+21.7) = 23.4 kg/mm2 "average" magnitude of 1/2(21.7+ 18.3) =- 20.0 kg/mm2 "average" magnitude of 1/2(18.3+14.9) = 16.6 kg/mm2

-=

=

There are 9,000 stresses with an "average" magnitude of 1/2(14.9+11.7) = 13.45 kg/mm2 There are 90,000 stresses with an "average" magnitude of 1/2(11.7+ 8.1) = 9.9 kg/mm2 There are 900,000 stresses with an "average" magnitude of 1/2( 8.1+4.8 ) = 6.45 kg/mm2 There are 9,000,000 stresses with an "average" magnitude of 112( 4.8+1.8 ) = 3.3 kg/mm2

These 7 fatigue loads replace the whole load

spec-trum of the "Canada". It is very difficult to find

a combination of one number of cycles and one stress which can be considered to be equivalent to these 7 loads.

It is even quite impossible to find a sufficient number of these combinations that a line can be

drawn of which each point represents a load which

is equivalent to the whole load history of the

"Canada" (fatigue loading line). Therefore we

only will give a line of which safely can be assumed

that it does not overestimate the loading of the

"Canada". The following values are used: 10 cycles of 23.4 kg/mm2

100 cycles of 20.0 kg/mm2 1000 cycles of 16.6 kg/mm2

etc.

In figure 5 these loads have been indicated by

"Deck-fatigue loading line"; in figure 7 a similar line is given for the bottom.

In section 9 these lines will be compared with Wohler curves for welded details.

9 The fatigue strength of the "Canada"

It has already been said (section 6) that although the mean value of the wave bending stresses is of

primary importance for the fatigue strength of a ship this influence cannot be shown in the

cu-mulative frequency distribution. It can, however,

be taken into account in the fatigue strength or Wohler curves with which the "fatigue loading line" of the "Canada" will be compared. For the

deck of the "Canada" we will use Wohler curves for repeating tensile loads because the still water stresses of this ship are tensile in the fully loaded condition.

In the bottom of the "Canada" the longitudinal still water stress, which is compressive together with the local stresses due to static waterpressure (see figure 11) will lead to large compressive stresses

-I- TENSION BULKHEAD OR FLOOR

COMPRESSION

BOTTOM LONGITUDINAL

+++++++++

wATERL.REssuRE

Figure 11

at certain points of the bottom and small

pressive or tensile stresses at other points. A com-pressive state of stress is generally favourable where fatigue is concerned.

So the dangerous points in the bottom are the

ones where rather a neutral state of static stresses is present. There the wave-bending and fluctuating

local stresses possibly combine into alternating

stresses. As a consequence the WOhler curves with which the bottom fatigue loading line will be com-pared must be curves for alternating loads.

A more accurate estimation of the most suitable level of the mean stress of the Wohler curves has not much sense, for the still water stresses in a ship

change as often as the loading condition or the temperature changes. Also the static part of the

local stresses differs from place to place as a

con-sequence of the presence of cargo, ballast, fuel,

engines, auxiliaries, the weight of the structure and the unavoidable residual stresses.

The fatigue strength curves for various welded

details in figures 12 and 13 are thought by the

author to be typical for shipbuilding. Some of them

refer to the structures of figure 3. In ships, butt

welds in deck and bottom will at least conform to

the ones indicated by "unmachined with small

weld faults". Figures 12 and 13 show clearly that these welds will be safe from fatigue if stress con-centrations are not present.

The lowest Miller curve in the diagrams applies to interconnections of longitudinal frames of the

types used in T2-tankers which are tested in the

Ship Structures Laboratory of the Technological University Delft. Use has also been made of data from [20].

The specimens at Delft are full size; they have been constructed by Wilton-Feyenoord at Schie-dam. They form part of an extensive investigation of the fatigue strength and brittle fracture strength

of large structural components. The load of the

specimens is somewhere between repeating tension and alternating loading. Therefore in figure 12 the

results have been modified in order to get a line for pure repeated loading. The author has used

an approximation:

1/2(2A + B) = peak to zero value of equivalent repeated load.

A is the tensile part of the dynamic load and B is the compressive part. For figure 13 use is made of:

1/3(2A B) = peak to zero value of pure

alternating load.

For our purpose these approximations are

suffi-ciently reliable.

The rather bad strength of the longitudinals is partly due to the fact that the neutral axis at the section adjacent to the transverse bulkhead is in

a higher position than elsewhere, which results in bending.

DOUBLE AMPLITUDE

OF STRESS (0-KG/mm 2

4

-From figure 12 can be deduced that in the deck

longitudinals of the "Canada" fatigue cracks will initiate after approx. 10 years if these connections are of the T2-tanker type. The bottom longitudinals seem to be safe from fatigue (see figure 13). Nev-ertheless this type of structure shouldbe avoided ; the classification societies already demand unin-terrupted structures for ships with a length greater than 230 m. For smaller ships the interrupted types

35 _ _ _ _ 30 _ ... _

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INrTIAL CRACK COMPLETE FAILURE 25

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Min/

I 10 102 103 104 105 106 107 108 N. Figure 12 35 34 33 32 31 30 29 28 27 26 25 24 23 22 21 20 19 18 17 16 15 14 13 12 11 10 9 8 7 6 5 3 2 0 ORIGINALT2 -CNNECTION

DOUBLE AMPLITUDE OF STRESS() 1KG/mm2 40 39 38 37 36 35 34 32 311 30 29 28 27 26 25 24 23 22 20 19 18 16 15 14 15 4 -, 3 2 1 . 0

are often preferred in shipyards, because of the

advantage during construction.

An idea about the difference in fatigue strength between interrupted and continuous longitudinals can be obtained in figures 12 and 13 if the curve

Figure 13

,of the T2-tanker longitudinals is compared with the curves for unmachined butts welded in difficult position.

It can also be seen that complete fatigue frac-tures are not likely to occur in ships because the

--, N. /... N

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\41ALL CRACK A COMPLETE FAILURE ---,....

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10. 102 103 104 106 107 '10f8 N. 33 21 17 13 12 11 10

number of cycles for the development of a large

crack is many times larger than the number of

cycles necessary for the initiation of a crack. So far fatigue has seemed to be of little

impor-tance in ships. The picture however is not yet

complete for 3 subjects must still be considered: Large openings and other large discontinuities in ships.

Corrosion fatigue.

The influence of fatigue on the brittle fracture strength.

The danger of large openings in ships is

well-known. At hatchcorners, stress

concen-trations can easily amount to 3; a value of 2.5

being quite normal. A corresponding fatigue line

is not given in figures 12 and 13 because the

"effective stress concentration" (with regard to

fatigue) is always smaller than the theoretical stress concentration, particularly if the stresses are high.

In any case in the "Canada" fatigue cracks are likely to occur after a few years if the hatch corners are constructed without special care.

Similar observations apply to other structural discontinuities like ends of superstructures and deckhouses.

Corrosion fatigue is a big problem for ship-owners, but also for scientists who have to face a field with many variables (corrosive types of cargo, protective layers, etc.). One thing is cer-tain: there is a large difference between fatigue in a neutral atmosphere and in a corrosive atmosphere as has already been shown in figure 1.

Ships are always in a corrosive environment

being either in fresh water or seawater. The cargo of tankers is often very aggressive. It depends on

the nature and quality of the protection against

corrosion how a structure will behave under fatigue loads. A good protection can keep up the fatigue strength to a high level as has been demonstrated in a long-range investigation (as yet unpublished)

with welded full scale girders in seawater in the

Stevin Laboratory of the Technological University at Delft.

In ships the surface protection is often worst at

places where structural elements are welded

to-gether because of bad accessibility. At the same

time high stress concentrations can be present

there. As a consequence local plastic deformations

due to relief of residual stresses and alternating plastic strains, can weaken the protective layer.

On account of these considerations it can be

con-cluded that corrosion fatigue will be present in many ships. The reduction in fatigue strength of

50% applied by Dr. YUILLE in figure 1 will often be too optimistic because it refers to high frequency loading. The low frequency of the loading of ships constitutes a very unfavourable factor.

The influence of fatigue on the brittle frac-ture strength of strucfrac-tures. So far fatigue has mainly been discussed as a phenomenon that can

result into initiation of cracks. It has been stated

that small fatigue cracks in ships generally will not easily develop into large failures because the life

of a ship is not long enough. Also if a crack of appreciable length should develop, it would be

discovered in time and repaired.

In addition to the initiation and propagation of cracks, another aspect of fatigue is of importance which often escapes attention viz, the damage in-duced in a material by fatigue loading before cracks are initiated. This fatigue damage can result in a

rise of the brittle fracture transition temperature

of shipbuilding steels. Tests in the Ship Structures Laboratory proved that for the Charpy-V-test, the 15 ft.lb transition temperature can often increase about 10 °C as a consequence of fatigue loading. The same applies to the 50% fracture appearance-transition temperature.

In structures in which small fatigue cracks are present it can be expected that the brittle fracture characteristics will be impaired to an even larger

extent. Nevertheless the fatigue damage of the

material outside the crack may be equally impor-tant. The first tests with full scale interconnections

of longitudinals of the T2-tanker type strongly

support this opinion.

10 Conclusions and final observations

In relatively fast dry cargo ships the wave

bending and slamming stresses are appreciably higher than in ships of moderate speed.

The longitudinal structure of ships of the

"Canada" type can suffer from high cycle fa-tigue as well as from low cycle fafa-tigue. High cycle fatigue is possible at minor discontinuities

such as interconnections of interrupted

Ion-gitudinals.

Low cycle fatigue can manifest itself at large discontinuities where appreciable stress con-centrations occur and welds of insufficient qual-ity are present.

With good structural design and sound work-manship there is little danger of fatigue in dry

cargo ships as long as the protection against

corrosion is "sufficient". Otherwise

corrosion-1..

ad 2.

fatigue cracks will develop easily because of

the low fatigue strength of unprotected struc-tures in a corrosive environment.

4. In ships of the "Canada" type the danger of brittle fracture should not be underestimated. The stresses in the deck can become very high

due to large still water stresses in the loaded

condition combined with high wave-bending-and slamming stresses.

The relatively severe dynamic loading

contrib-utes to the risk of brittle fracture in another

way too, because a deterioration of the material can be caused and/or minor cracks may develop. When in the future, larger ships are to be built or the speeds are increased (atomic propulsion), im-proved strength calculations will become a neces-sity. Nowadays the attention in ship research seems

to be particularly directed to the extreme loads

which ships will meet in their life. The danger of

brittle fracture or collapse due to buckling are considered of primary importance. On the other

hand many people are convinced that the problem of brittle fracture has been overcome by the elim-ination of abrupt discontinuities and the applica-tion of improved steels. This point of view is not quite correct for if higher stresses would be allowed

in ships the problem could rise again. In the

author's opinion, however, it is quite possible to

use steels of such quality that brittle fracture will not occur as long as the nominal stresses are below yield point. This can hold even if minor

fatigue-or weld cracks are present. Consequently there

should be no objections against higher longitudinal

stresses in ships than is average practice now if

only brittle fracture is concerned.

The next problem is buckling. Here higher

lon-gitudinal stresses will be possible if and when a

more efficient distribution of the longitudinal and

transverse material in ships is realised.

Theoret-ically, solutions of this problem can be found and the time will come that they have to be accepted in practice.

After that, the admissible longitudinal stresses will mostly be governed by fatigue criteria. This can be illustrated in the following way. In figure 5

it can be seen that the highest double amplitude of stress in the deck of the "Canada" is equal to

25.1 kg/mm2. From figure 12 we have concluded that the complete load spectrum of the "Canada" is rather severe from a viewpoint of fatigue strength.

In other words in that respect the longitudinal

structure of the "Canada" is not too strong. We will now consider the strength with regard

to the extreme loads in this ship. The maximum

value of the still water stresses (hogging) is about 6 kg/mm2 [12] ; the maximum amplitude of the fluctuating stresses amounts to 25.1 : 2 = 12.5

kgimm2.1) Together they give rise to a maximum stress in the deck equal to 6+12.5 = 18.5 kg/mm2.

It is difficult to say how accurate this value is.

Probably it is not lower than in reality because it

has mainly been obtained by straight-line

extra-polation of experimental results (figures 4 and 5) which can lead to excessive extreme values for fast dry-cargo ships [24].

The value 18.5 kg/mm2 is rather high for present-day conceptions, but in relation to structural strength up to yield point there is a definite margin of safety.

It will be clear that when in ships the still water stresses are low the margin of safety as regards

brittle fracture and compressive instability is even appreciably higher. In that case the extreme stress in the deck should even be restricted to 15 kg/mm2 in stead of 18.5 kg/mm2 if fatigue-cracks are to be avoided.

Summarizing it can be said that where structural design permits static stresses close to yield point or when still water stresses can be kept low through-out a ship's life, the longitudinal strength of a ship will mainly be governed by fatigue criteria. In that case the currently interesting problem concerning

the use of high-strength steels in ships looses much of its importance for the fatigue properties of weld-ed structures made of special steels are not signif-icantly better than those of mild steel structures. In fact steels possessing high yield strength are only

usefull when the maximum of the nominal

lon-gitudinal stresses is allowed to be higher than yield

point of mild steel. This situation will only be

possible when the wave bending moments are rel-atively small and the still-water bending moment

is very large. In that case fatigue strength con-stitutes no problem and the design can be based

on the magnitude of the extreme loads only. Even

then it is doubtfull if compressive stresses up to

the yield point of high strength steel can ever be

permitted, for in present-day practice the elastic

stability of most ship structures does not even per-mit stresses up to the yield point of mild steel.

So far the shape and inclination of the cumu-lative frequency distribution of the longitudinal stresses has not received much attention in this

section. This necessitates some reserve as to the applicability of the conclusions to other types of ships than fast dry cargo-ships. This applies

partic-ularly to tankers in which the extreme values of ') The hogging part of the wave bending stresses has

the wave bending stresses are higher than in dry cargo ships [24].

When fatigue considerations should become a leading factor in ship structural design the loading for each ship must be known before the design can be commenced.

A possible method for a reliable estimation of

this problem is the use of model tests in waves. The feasibility of these tests has been sufficiently proved [19]. This type of modeltesting could become

stan-dard procedure for ships together with the usual

tests on seakindness, resistance and propulsion. The knowledge about the behaviour of the sea-surface seems to be sufficiently advanced now to realise the proposed proceedings.

References

1 I.S.S.C. report of the Committee on Response to Wave

Loads, (D.T.M.B. report no. 1537), June 1961.

YUILLE, I. M.: Longitudinal Strength of Ships. R.I.N.A.

1962.

JOURDAIN, M.: Premiers resultats de l'etude statistique

des contraintes a. la mer sur des navires de commerce,

A.T.M.A. 1961.

VEDELER, G.: To what extent do brittle fracture and

fatigue interest shipbuilders today, Houdremont

lecture 1962, Sveiseteknikk 1962; no. 3.

STENEROTH, E.: Low cycle fatigue, I.S.S.C. 1961.

Gisns, H. R., and G. M. Boy: Welded Ship Construc-tion - Record of common Fractures and their causes,

Trans. Inst. of Eng. & Shipb. in Scotland; Part 5,

1957-'58.

Reports on Fatigue Failures, Comm. XIII of I.I.W.,

Report I.I.W./I.I.S. 14-59, Brit. W. J. 1961.

ABRAHAMSEN, E.: Tank size and dynamic loads on

bulk-heads in tankers, Europ. Shipb. no. 1; 1962.

YOSHIKI, M.; Y. YAMAMOTO and K. FIAGIWARA :

Ex-periments on dynamic pressure in cargo-oil tanks

due to ship motions. J. Soc. Nay. Arch. Japan 109,

B.S.R.A. 17.992.

WATANABE, Y.: On the water pressure in the tank due

to rolling of a ship. Kyushu Univ. Fac. of Eng. vol. 16 no. 4, 1957, B.S.R.A. 13,404.

BENNET, R.: Stress and motion measurements on ships

at sea; part I-II, Rep. no. 13; 1958, The Swedish

Shipb. Res. Ass.

Idem part III, Rep. no. 15; 1959.

It should be remembered, however, that load

spectra of ships in any form cannot yet be checked

against fatigue strength data of structural details

loaded in a similar way. Programmed fatigue test-ing of structural components, preferably in corro-sive atmosphere, seem to be far ahead. The testing which will be done in the near future will be con-centrated on high and low cycle fatigue problems with comparable structures in order to obtain

rel-ative fatigue data. This is quite natural because

the need for this information is most urgent. Nevertheless the above-mentioned model testing should be stimulated because it may provide more reliable information for the structural design of a ship than any statical method can do.

KORVIN KROUKOWSKI, B. V.: Ships at sea, (Chapter V)

Stevens Institute of Technology, Jan. 1958.

AERTSSEN, G.: Service-performance and seakeeping

trials on m.v. Lukuga, T.R.I.N.A., March '63.

WARNSINCK, W. H., and M. ST. DENIS: Dutch destroyer

trials, Proc. of Symp. on the behaviour of Ships in a seaway vol. 1; 1957.

BLEDSOE, M. D.; 0. BUSSEMAKER and W. E. CUMMINS:

Seakeeping trials on three Dutch destroyers.

JASPER, N. H.: Service stresses and motions of the Esso

Asheville, T.M.B. rep. 960; Sept. '55.

JAEGER, H. E., and J. J. W. NIBBERING: Beam knees and other bracketed connections, I.S.P. Jan. 1961. DOES, J. Ch. DE: Experimental determination of bend-ing moments for three models of different fullness in regular waves, I.S.P., Feb. '60.

LEIRIS, H. DE, and H. DUTILLEUL : Etude comparee de

quelques assemblages de coque, facilite de montage et concentrations de tension, A.T.M.A. 1951. WECK, R.: Fatigue in shipstructures, R.I.N.A. 26 Mrch

1953.

92. NEUMANN, A.; H. POHL; G. MOLLER and H. KATFINER :

Bauteiluntersuchungen fur den Schiffbau, Schweiss-technik, August 1955.

JACKWITZ, H., and G. MULLER :

Dauerfestigkeitsunter-suchungen an Konstruktionsverbindungen von

Bodenwrangen mit Langsbandern, Schweisstechnik,

April 1958.

BENNET, R.; A. IVARSON and N. NORDENSTROM: Results

from full scale Measurements and predictions of wave bending moments acting on ships,

Skepps-byggnadsteknisk Forskning Report no. 32, 1962.

13. 14, 15. 17. : 19., 20L : 2.

-No.. S

No. 2 No. .3S

No. 4.S

No. 5S

No.. 6 S Some tests on stayed and unstayed masts and a comparison of experimental results and calculated stresses' (Iltitch).

By ir A. Verduin and ir B. Burglkgraef. June 1952.

No 7 M. Cylinder wear in marine diesel engines (Dutch).

By ir H. Visser. December 1952.

No. a M Analysis and testing of lubricating oils (Dutch).

By ir R. N. M. A. Malotaux and ir j. G. Smit. Jury 1953.

Nq., 9 S ';Stability experiments on models of Dutch and French standardized lifeboats.

By prof ir H. E. Jaeger, prof. ir J. W. Bonebakker and J. Pereboom, in collaboration with A. Andige. October 1952.

No. 10 S On collecting ship service performance data and their analysis.

By prof. ir j. W. Bonebakker. January 1953.

No,.. 11 M The use of three-phase current for auxiliary purposes (Dutch).

By ir j. C. G. van Wijk. May 1953.

No. 12 M Noise and noise abatement in marine engine rooms (Dutch).

By "Technisch-Physische Dienst T.N.0.-T.H." April 1953.

No. 13 M Investigation of cylinder wear in diesel engines, by means of laboratory machines (Dutch).

By ir H. Visser. December 1954.

No. i4 .114 The purification of heavy fuel oil for,diesel engines (Dutch).

By A. Bremer. August 1953.

No.. 1Lr S Investigation of the stress distribution in corrugated bulkheads with vertical troughs.

By prof. ir H. E. Jaeger, ir B. Burghgraef and I. van der Ham. September 1954.

No. 16 M Analysis and testing of lubricating oils II (Dutch).

By ir R. N. M. A. Malotaux and drs j. B. Zabel. March 1956.

No. 17 M The application of new physical methods in the examination of lubricating oils,.

By ir R. N. M. A. Malotaux and dr F. van Zeggeren. March 1957.

No. la m No.. 28 Nit No. 29 M No. 30 S No. 31 M Reports

The determination of the natural frequencies of ship vibrations (Dutch),..

By prof. ir H. E. Jaeger. May 1950.

Confidential report, not published. July 1950.

Practical possibilities of constructional applications of aluminium alloys to ship construction.

By prof. ir H. E. Jaeger. March 1951.

Corrugation of bottom shell plating in ships with all-welded or partially welded bottoms (Dutch),

By prof,itH. E. Jaeger and ir H. A. Verbeek. November 1951.

Standard-recommendations for measured mile and endurance trials of sea-going ships (Dutch..

By prof ir J. W. Bonebakker, dr ir W. J. Muller and ir E. J.Diehl. February 1952.

Considerations on the application of three phase current on board ships for auxiliary purposes especially with regard to fault protection, with a survey of winch drives recently applied on board of these ships and their

in-fluence on the generating capacity (Dutch).

By ir y. C. G. van Wijk. February 1957.

No. 19 M Crankcase explosions (Dutch).

By is_J.H. Minklwrst. April 1957.

No. 20 S An analysis of the application of aluminium alloys in ships' structures.

Suggestions about the riveting between steel and aluminium alloy ships' structures.

By pro f. itH. E. Jaeger. January 1955.

No. 21 S On stress calculations in helicoidal shells and propeller blades..

By dr ir_J. W. Cohen. July 1955.

No. 22 S Some notes on the calculation of pitching and heaving in longitudinal wa.ves.

By ir 3.Gerritsma. December 1955.

No. 23 S Second series of stability experiments on models of lifeboats,. By ir B. Burghgraef. September 1956.

No. 24 M Outside corrosion of and slagformation !on tubes in ,oit-fired boilers '(Dutch).

By dr W. 3. Taal. April 1957.

No. 25 S 'Experimental determination of damping, added mass and added mass moment of inertia of a shipmode .

By it J. Gerritsma. October 1957.

No. 26 M Noise measurements and noise reduction in ships.

By ir G. j. van Os and B. van Steenbrugge. May 1957.

No. 27 S Initial metacentric height of small seagoing ships and the inaccuracy and unreliability of calculated curves of righting levers.

By Prof ir 3 W. Bonebakker. December 1957.

Influence of piston temperature on piston fouling and piston-ring wear in diesel engines using residual fuels.,

By ir H. Visser. June 1959.

The influence of hysteresis on the value of the modulus of rigidity of steel.,

By Zr A. Hoppe and ir M., Hens. December 1959.

An experimental analysis of shipmotions in longitudinal regular waves.

By ir j. Gerritsma. December 1958.

1

h

Model tests concerning damping coefficients and the increase on ship's propellers.

By N.J. Visser. October 1959.

i

in the moments of inertia due to entraine water.

-(Continuation of report No. 17 M.)

By ir R. N. M. A. Malotaux and dr F. van Zeggeren. November 1959.

No. 34 S Acoustical principles in ship design. By ir 3. H. Janssen. October 1959.

No. 35 S Shipmotions in longitudinal waves.

By ir 3. Gerritsma. February 1960.

No. 36 S Experimental determination of bending moments for three models of different fullness in regular waves.

By ir 3. Ch. De Does. April 1960.

No. 37 M Propeller excited vibratory forces in the shaft of a single screw tanker.

By dr ir 3. D. van Manen and ir R. Wereldsma. June 1960. No. 38 S Beamknees and other bracketed connections.

By prof. ir H. E. Jaeger and ir 3. 3. W. Nibbering. January 1961

No. 39 M Crankshaft coupled free torsional-axial vibrations of a ship's propulsion system. By Jr D. van Dort and N. 3. Visser, June 1963.

No. 40 S On the longitudinal reduction factor for the added mass of vibrating ships with rectangular cross-section. By ir W. P. A. Joosen and dr 3. A. Sparenberg. April 1961.

No. 41 S Stresses in flat propeller blade models determined by the moire-method.

By ir F. K. Ligtenberg. June 1962.

No. 42 S Application of modern digital computers in naval-architecture.

By ir H. 3. Zunderdorp. June 1962.

No. 43 C Raft trials and ships' trials with some underwater paint systems.

By drs P. de Wolf and A. M. van Londen. July 1962.

No. 44 S Some acoustical properties of ships with respect to noise-control. Part I. By ir 3. H. Janssen. August 1962.

No. 45 S Some acoustical properties of ships with respect to noise-control. Part II. By ir 3. H. Janssen. August 1962.

No. 46 C An investigation into the influence of the method of application on the behaviour of anti-corrosive paint systems in seawater.

By A. M. van Londen. August 1962.

No. 47 C Results of an inquiry into the condition of ships' hulls in relation to fouling and corrosion.

By ir H. C. Ekama, A. M. van Londen and drs. P. de Wolf. December 1962.

No. 49 S Distribution of damping and added mass along the length of a shipmodel.

By prof. ir 3. Gerritsma and W. Beukelman. March 1963.

No. 50 S The influence of a bulbous bow on the motions and the propulsion in longitudinal waves.

By prof. ir 3. Gerritsma and W. Beukelman. April 1963.

No. 52 C Comparative investigations on the surface preparation of shipbuilding steel by using wheel-abrators and the application of shop-coats.

By ir H. C. Ekama, A. M. van Londen and ir 3. Remmelts. July 1963.

No. 53 S The braking of large vessels.

By Prof. ir H. E. Jaeger.

No. 54 C A studit of ship bottom paints in particular pertaining to the behaviour and action of anti-foliling paints.

By'A. M. van Londen. September 1963.

No. 55 S Fatigue of ship structures.

By ir 3.3. W. Nibbering. September 1963.

Communications

No. 1 M Report on the use of heavy fuel oil in the tanker "Auricula" of the Anglo-Saxon Petroleum Company (Dutch). August 1950.

No. 2 S Ship speeds over the measured mile (Dutch).

By ir W. H. C. E. Roiingh. February 1951.

No. 3 S On voyage logs of sea-going ships and their analysis (Dutch).

By prof. ir 3. W. Bonebakker and ir 3. Gerritsma. November 1952.

No. 4 S Analysis of model experiments, trial and service performance data of a single-screw tanker. By prof. ir J. W. Bonebakker. October 1954.

No. 5 S Determination of the dimensions of panels subjected to water pressure only or to a combination of water pressure and edge compression (Dutch).

By prof. ir H. E. Jaeger. November 1954.

No. 6 S Approximative calculation of the effect of free surfaces on transverse stability (Dutch). By ir L. P. Herfst. April 1956.

No. 7 S On the calculation of stresses in a stayed mast.

By ir B. Burghgraef. August 1956.

No. 8 S Simply supported rectangular plates subjected to the combined action of a uniformly distributed lateral load and

compressive forces in the middle plane. By ir B. Burghgraef. February 1958.

No. 9 C Review of the investigations into the prevention of corrosion and fouling of ships' hulls (Dutch). By ir H. C. Ekama. October 1962.

M engineering department

S = shipbuilding department

C = corrosion and antifouling department =

=