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
VERMOEHNG VAN SGHEEPSCONSTRUCTIES
by
IR. j.j. W. NIBBERING
(Senior-scientific_officer of the Ship Structures Laboratory Drift)
Issued by the Council
This report is not to be published unless verbatim and unabridged
REPORT No. 55S
SSL 94a
September 1963CONTENTS
page
Summary 3
i Introduction 3
2 General opinions, practical experience and results of
fatigue tests 3
3 Dynamic loading of the longitudinal structure of ships 5
4 Slamming ...7
5 Corrosion allowánces and tdlerances of thickness . IO
6 The influence of variations of temperature and changes in
loading condition . . 10
7 Local bending of longitudinal members of the bottom due
to waterpressure i 12
'ft Actual peak to peak values; fatigue loading line . 12
9 The fatigue strength of the "Canada" . 15
IO Conclusion and final observations 18
i 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 inditi1orthedesira-bi1it-y-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, practicalexperience 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 practicediffers 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
FATIGUE OF SHIP STRUCTURES*)
byIR. J.J. W. NIBBERING
Summary
Fatigue is often considered to be a problem of secondary importance in naval architecture. The correctness óf this opinion will be investigated in this article For this purpose the loading of the longitudinal stiucture 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
*) Publication no. 12 of the Ship Structures Laboratory of the Technological University, Delft
1) For convenience, this value will be called "stress" hereafter.
3
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 aftér 30 years [1]. But ¡t must be mentioned that the corresponding ln 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 krown "Ocean Vulcan".
Iflnktthe-line--3O--years-infigureJ_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 spécial curve for the stresses at
dis-continuitieswith a stress concentration factor equal to 3. This curve nearly touches the fatigue-strength curye for notched specimens in corrosivé
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.
J: France JOURDAIN [3] has come to a similar
conclusion. In practice many fatigue cracks have been found. VEDELER for instance does not doubt
0.35
that most of the cracks found in ships re due to
fatigue [4]Ç STENEROTH [5]. is 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 ôf 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/o yeaÈs. In figure .2
FATIGUE LIFE
- PLAIN MS SPECIMENS
_FAflGUE LIFE
PLATES -WITH BUTT 'WELDS CONTAINING FLAWS
\
\
.-.
TYPICAL FATIGUE LIFE
UNDER CORROSIVE CONDITIONS
---.--.-_-
N
10 102 io NUMBER OF CYCLESFigure 1 Diagram according to Dr. Yuille [2]
N
100 o 1O
(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 cavgo- 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
concentratiòn .at its origin. But the fact that the
4 0.25 od
z
I-02 4000-u, u' w I-(j w, a-z 4 I-di) t 3000 -w z 4 o w z 4 0.15 I. -J -J. 4 2000- I- 01 1000 005 4 7000 6000 -w I. -O-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.
Figure 2
The foregoing discussion seems to indicate that the dynamic loading Of :bulkli'äds in tankers de-serves much more attentiön 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 atteñtion 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, JAcKwI.Tz, MÜLLER
and others worked in the same fleicLand also tested bilge frame connections and connections of contin-uous longitudinal: frames to transverse webs and bulkheads [22] [23]. De. LEIRi5 and DLJTILLEUL investigated interconnectionsof longitudinal frames at tranverse bulkheads [20]. Finally JAEGER and
NIBBERING have tested various connections :between
orthogonally placed stiffeners like beämknees [18].
In the latter investigation many specimens were
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 iii 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
SMAll. CRACK
FA11ÍGUE BENDING TESTS [L]
E. E. 10
o-I
45 15ÌIkUIlUUEIIiiliII
II!!HIIIii!!!!iiI
11111 iIIIIIIIIIIIIIIiuiiuuiiiiiui
io' io4 io' 106 iO' NUMBER OF.CVCLES-said that the dynamic loading of a ship in seaway does not seemtobe-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 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.
12
lo io2 io l0 lo
N=NUMBER OF TIMES THAT THE INDICATED STRESSES HAVE BEEN EXCEEDED.
Figure 4
THE POSIrIONOF THE STRUCTURES IHORIZON'TALLVI REFERS ONLY TO THE
IFA1IGUE LIFE OF THE DETAILS
INDICATED BY ARROWSn2x103 n9x103 n25x1O4 6.5xTO4 I 1x103 1x104 1x105 1x10'6 1x107 Figure 3 DOUBLE AMPUTUDE OF STRESS() KG/mm2 106 l0 108 N 25 25 24 23 -'22 21 T 20 20 19-18 - HORIZONTALLY WELDED 17 - U NMACHIN E D.y I 16-15
15...
14 -lo 1longitudinal 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 i month seems to be as severe as the löadspectrurn of the "Ocean Vulcan" for 30 years. This must be traced to the differences between both ships in speed (7 knots), dimensions, form ançi type. The blockcoefficient of the "Ocean Vulcan" was 0.75 the "Canada" had 0.65. The "Ocean Vulcan" is' a full scantling ship. The "Minflesota" 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 streses 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 longiiiidiiiäl bëii&ng stresses. These corrections are related to:
Slamming and whipping;.
Changes.in temper'ature and loadingconditions;
e. Changes in waterpressure on the bottom in
seaway;
d. Deviations in plate thickness due to orrosion 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 stùesses should' be taken into accountwhen 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 seconchsquite 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 brQader sense than is usual. Any load causing vertical two nçde 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.
Korzin-Kroukowski ([13]. pages V36, V40, V4 1)
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 slammedduring 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. Slamming of fast ships:
"The 'submergence of the bow and the dynamic effect of t'he 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 anec-essary prerequisite'to the occurrence.of slam. . ". . .In additiOn to the severe shocks recorded
as slams, the frequent bow immersion in the head seas caúses shocks of sufficient 'magnitude
to maintain the hull in a continuous state of vibration."
7
a. lo
8
b'. R. Rennet ([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 thesame size." ". .. No account is taken of the biggest slam-ming peaks, but only of the normally occurring slamming."
"The maximum value includes frequently oc-cuiring 'slamming increases but not the occa-sional, very big peaks which were observed particularly on the Minnesota."
c. 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 UulF'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 l0-30seconds.
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 supeiiposed on the stress curves and these oscillations - of a stress range
of 0.8 to 2 tons/sq.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 spèed. 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 Hl/lo = 18 ft. In light-load condition, how-ever, even in a sea about 1' 1 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 ofthe longitud-mal 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, butthat the results have not (yet) been published.
The present article is not meant to cover a,
thOrough analysis of'slamming 'in ord'er 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:
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.
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 tò 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
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 15 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 sfress.. (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 slám 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 tha t 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, atordingtoTable-Ionly--3-slams-would-have-oe---curred in stead of 7. Along the same line it can be
TABLE I Number of slams pçr 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:
The speed of the "Canada' is as high as 19.5 knots.
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 comparablewith the destroyers.
Generally the destroyers were fully-loaded while
during 45% of the time the "Canada" sued 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 testsof 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/mm2 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 T'i have been derived from the frequency distribution line of the "Canada" for I month in figure 4. In column (5) the number of slams of the intensities of column (2) have-been-estimated._Theseslanirstresses are given in proportion to the maximum slam-stressthat
prob-.9
Speed Heading
Ship A Ship B ' Ship S
Slams encounterCycles of Slams encounterCycles of Slams encounterCycles of
12 knots Head 7 120 9 95' 3 97
12 knots Bow IO 110 2 105 7 97
12 knots Beam 4 105 2 92 2 89
12 knots Quartering 6 , 93 3 87 0 75
12 knots Following 0 60 1 47 0 62
17 knots Head 19 102 19 112 3 loo
17 knots Bow 14 112 9 112 16 100
17 knots Beam 5 103 7 97 7 100
17 knots Quartering 0 94 0 96 0 87
17 knots Following 0 71 0 70 0 62
lo
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' 111e 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 i = 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)
()
(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 ofcycles for Estimated wave stresses number of
larger than . . slams indicated in Total number of slams Indication in figure 5
(la) (lb) column (la)
0.95 à S 0.9 x 0.5S
-
.-
-0.7x0.5S 1.255 j . 1 1 0.85 à 0.9S 0.9x 0.5S l.25S 1 2 1 0.7x0.55 l.15S 3 5 2 0.5 x 05S 1.055 3 . 8 0.7S à 0.85 0.9x 0.55-
.-
-0.7x0.5S 1.055 4 12 3 0.5x0.5S 0.955 . 8 20 0.3x0.5S 085S u-, . 6 26 0.6S à 0.75 0.9 x 0.55-
. -0.7x0.5S 0.95S 12 38 4 0.5x 0.5S 0.855 30 68 5 0.3x0.5S 0.755 30 98 0.5S à 0.6S 0.9 x 0.55-
.-
-0.7x0.5S - . . .-
-0.5x05S .0.75S . 50 148 6 0.3x0.5S 0.655 70 218 7 0.1 x 05S 0.555 80 298 0.35 à 0.55 0.9x 0.55 .-
.-
-0.7x0.5S --
-0.5x0.5S 0.555 52 350 8 0.3x0.55 0.455 500 850 . 9 0.lxO.5S 0.355 1,000 1,850 10 0.lSàO.3S *-0.lxO.55 0.l5S 2,650 4,500 lIDOUBLE AMPLITUDE.
OF STRESS() KG/mm2
N
NNUMBER OF TIMES THAT THE INDICATED STRESSES HAVE BEEN EXCEEDED.
Figure 5 Fatigue loading of deck
15x104 io'
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.
35 30 25 20 15 10 5 O
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/mma. 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 be 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
11 HORIZONTALLY WELDED; 10. io2 io8 Ñ. 35 34 33 32 31 -30 29 28 27 26 -25 24 23 - 2221 -20 19 18 17 16 -15 14 13. 12 11 -10 9- 87. - 6-5 4 3 2 O
12
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/mm2. 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 a 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 8kg/mm2; 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).. Mòreover, 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 figire 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 div.ided 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: (H H2)m
waterpressur.e).
e. Variations in the pressure of the cargo on the bottom during heaving, pitching and rolling as a consequence of the inertia of the cargo. d. Similar variations in waterpressure
(hydro-dynamic load).
These components are interrelated in a complex
way. Phase differences exist between them. Moreo-ver, each one isa 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 (e. and d. will counteract to a certain extent) will support this opinion sufficiently.
The draught of the "Canada" was 7.5 m. In
the stiiwater 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
DOUBLE AMPLITUDE OF STRESS(Q) KG/mm2 35 34 33 32 31 30 29 28 27 26 -25 24 - 2322 21 -20 19 18 17 16 -15 14 13 12 11 -10 g 8
N=NUMBER OF TIMES THAT THE INDICATED STRESSES HAVE BEEN EXCEEDED.
Figure 7' Fatigue loading of bottom structure
357
UÑMAK [i] Figure 10 35 30 25 20 15 io 5 O io8 N 13 '4."
% '%zc> -A' , %.q;\'e
%\ % ..". .., fr..
(, %\
eor*
.\
10 io2 io4 10i7
14
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 O to 7 comprises 7 half cycles of the following stresses:
O to 1 3.30 kg/mm2 i 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/mm2
If we now consider the peak to peak value between the lines O 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 O 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 öl cycles the larger this percentage can be. On the
average roughly i 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 structurai 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-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 no 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 iii a fatigue load of constant magnitude for a certain number of cycles. Such a rule does not exist. The well-known Miner's rule (En/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:
There are 9 stresses with an "average" magnitude of'/2(25.1+2l.7) = 23.4 kg/mm2 There are 90 stresses with an "average" magnitude of 1/2(21.7+ 18.3) = 20.0 kg/mm2 There are 900 stresses with an "average" magnitude of 1/2(18.3+ 14.9) = 16.6 kg/mm2
10 - i stresses having a magnitude between 25.1 and 21.7 kg/mm2 100 - lo 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 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
There are 9,000 stresses with an "average" magnitude of'/2(14.9+ 11.7)
=
13.45 kg/mm2These 7 fatigue loads replace the whole load spec- at certain points of the bottom and small corn-trum of the "Canada". . It is very difficult to find pressive or tensile stresses at other points. A corn-a combincorn-ation ofone number ofcycles corn-and one stress pressive state ofstress is generally favourable where which can be considered to be equivalent to these fatigue is concerned.
7 loads. So the dangerous points in the bottom are the
It is even quite impossible to find a sufficient ones where rather a neutral state of static stresses number of these combinations that a line can be is present. There the wave-bending and fluctuating drawn of which each point represents a load which local stresses possibly combine into alternating
is equivalent to the whole load history of the
stresses. As a consequence the Wöhler curves with"Canada" (fatigue loading line). Therefore we which the bottom fatigue loading line will be corn-only will give a line ofwhich safely can be assumed pared must be curves for alternating loads. that it does not overestimate the loading of the A more accurate estimation of the most suitable "Canada". The following values are used: level of the mean stress of the Wöhler curves has 10 cycles of 23.4 kg/mm2 not much sense, for the still water stresses in a ship 100 cycles of 20.0 kg/mm2 change as often as the loading condition or the
1000 cycles of 16.6 kg/mm2 temperature changes. Also the static part of the
etc. local stresses differs from place to place as a
con-In figure 5 these loads have been indicated by sequence of the presence of cargo, ballast, fuel, "Deck-fatigue loading line"; in figure 7 a similar engines, auxiliaries, the weight of the structure and line is given for the bottom, the unavoidable residual stresses.
In section 9 these lines will be compared with The fatigue strength curves for various welded Wähler curves for welded details, details in figures 12 and 13 are thought by the
author to be typical for shipbuilding. Some of them
9 The fatigue strength of the "Canada"
refer to the structures of figure 3. In ships, butt It has already been said (section 6) that although welds in deck and bottom will at least conform to the mean value of the wave bending stresses is of the ones indicated by "unmachined with small primary importance for the fatigue strength of a weld faults". Figures 12 and 13 show clearly that ship this influence cannot be shown in the cu- these welds will be safe from fatigue if stress con-mulative frequency distribution. It can, however, centrations are not present.be taken into account in the fatigue strength or The lowest Wöhler curve in the diagrams applies Wähler curves with which the "fatigue loading to interconnections of longitudinal frames of the
eof
Ship Structures Laboratory of the Technological University Delft. Use has also been made of data from [20].
The specimens at Deift 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:
'/2(2A+B) = peak to zero value of equivalent
amic load and B is the compressive part. For figure 13 use is made of: deck of the "Canada" we will use Wöhler 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
ROTTOM LONGITUDINAL WATERFRES5Jfi ROTTOM LONGITUDINAL WATERFRES5Jfi Figure 11 Figure 11 15 15
There are 90,000 stresses with an "average" magnitude of'/2(1.1.7+ 8.1) = 99 kg/mm2 There are 900,000 stresses with an "average" magnitude of '/2( 8.1 +4.8 ) = 6.45 kg/mm2 There are 9,000,000 stresses with an "average" magnitude of '/2( 4.8+ 1.8 ) = 3.3 kg/mm2
+ TENSION BULKHEAD OR FLOOR
16
I
'/a(2A+B) = peak to zero value of pure From figure 12. can be deduced that in the deckalternating load. longitudinals of the. "Canada" iàtigue cracks will For our purpose these approximations are suffi initiate after approx. 10 years if these connections
ciently reliable. areof the T2-tankertype. The bottom longitudinals
The rather bad strength of the longitudinals is seem to be safe from fatigue (see figure 13). Nev-partly 'due to the fact that the neutral axis at the, ertheless this type of structure shouldbe avoided; section adjacent to the transverse bulkhead is in the classification societies already demand unin-a higher position thunin-an elsewhere, which results in terrupted structures for ships with a length greater
bending, than 230 rn. For smaller ships the interrupted types
DOUBLE AMPLITUDE OF STRESS() KG/mm2 - 35 V V 30 O -. L
INITIAL CRAiI( COMPLETE FAILURE
HORIZONTALLY WELDED: MACHINED
---t=y=J---' 25 A -20-, - ' --'V',.
'
-. HORIZONTALLY_WELDED: LAIHACHINEDVIVI
z MODIF ED T2-CONNECTION p V i . V s ' UNMACHINED 15 T ÇVELDEO IN DIFFICULT POITTION . \'*1-ORIGIÑAIT2-CCNNECTION V -''-k
SEE FELT -TU Si i 10 io2 10 108 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 12il
10 9-8-7.
6-54.
3-
2-
1-
0-IlullIrili
ORIGINAL%T CON DOUBLE AMPLITUDE OF STRESS'(V) KG/mm2 Lo 39 38 37 36 35 34 33 32 31 30 29 28 27 26 25 24 4 23 22 18 17 16 15 14 13 12' 11 10 9 21 -20 19-
8-
6-5 4-
3-
2-O4 (9are 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'and13 if the curve
Figure 13 LO 35 30 25 20
of the T2tanker 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
17
18
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.
ad 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 beiñg 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 öbservations apply to other structural discontinuities like ends of superstructures and deckhouses.
ad 2. 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 atniosphere 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 Deift.
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. YULLLE in figure 1 will often be too optimistic because it refers to high frequency loading. The iow frequency of the loading of ships
constitutes a very urifavourable factor.
ad 3. 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 appreciäbly 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-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/orminor cracksmay 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 t 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 steçls 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 conçerned
The next problem is buckling. H'ere 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 oífatigue-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 kg/mm2.') 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 israther high for present-day conceptions, butin relation to structural strength up to yield point there is a definite marginof 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 t'he deck should even be restricted to 1.5 kg/mm2 in stead of 18.5 kg/mm2 if fatigue-cracks are to be avoided.
Summariing it can be sa'id' 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 óf 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 norriinal' lôn-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-ativay_smallnd_the_sti1lwater_bending moment is very large. In that case fatigue strength con-stitutes no problem and the 'design can be basçd on the magnitude of the extreme loads only. Even then it is doubtfull if' compressiv 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 necessi'tates 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
1) The hogging part of the wave bending stresses has
approximately been taken equal to the sagging part.
20
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 inwaves. 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.
J0URDAIN, M.: Premiers résultats de l'étude statistique des contraintes'à lamer sur desnavires 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.SC. 1961. GIBBS, H. R., and G. M. BOYD: Welded Ship
Construc-tion - Record of common Fractures and their causes, Trans. Inst. of Eng. & Shipb. in Scotland; Part 5
I 957-'58.
Reports 'on Fatigue Failures, Comm. XIII of I.I.W., Report I.I.W/I.I.S.14-59, Brit. Wj. 1961. ABRAHAMSEN, E.: Tank size and dynamk loads on
bulk-heads in tankers, Europ. Shipb no. 1; 1962.
YosHnu, M..; Y. YAMAMOTO and K. HAGIWARA: 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,, BSR.A. 13,404.
BENNET, R.: Stress and motion measurements on ships at sea; part J-II, Rep. no. 13; 1958, The Swedish Shipb. Res. Ass.
1'2 Idem part I'll, 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 KRouKowsK,I, 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.1.N.A.,, March '63.
WARNSINcK, W. H., and M. Sr. DENIs: Dutch destroyer
trials, Proc. of Symp. on the behaviour of Ships in aséaway vol.1; 1957.
l6 BLEDSOE, M D.; BUSSEMAKER and W. E CUMMIN5:
Seakeeping 'trials on threeDutch 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 '60k
20; LEIRI5, H. DE, and H. DUTILLEUL: Etude comparée de
quelques assemblages de coque, facilité de montage et concentrations de tension, A.T.M.A. 1951. 2h WEeK, R.: Fatigue inshipstructures, R.I.N.A. 26 Mrch
l953
NEUMANN, A.; H. POHL;' G. MÜLLER and H. KATHNER:
Bauteiluntersuchungen für den Schiffbau,
Schweiss-technik,. August 1955.
JACKWITZ, H., and G. MÜLLER:
Dauerfestigkeitsunter-suchungen an Konstruktionsverbindungen von
Bodenwrangen mit Längsbändern, Schweisstechnik, April 1958.
BENNET, R; A IvARsoN and N. NORDENSTRÖM':' Results
from full scale Measurements and predictions of wave bending moments acting on ships, Skepps-byggnadsteknisk Forskning Report no. 32, 1962.
No. 21 S No. 22'S No. 23 S No.24 M No. 25 S No.26 M No. 27 S
PUBLICATIONS OF THE NETHERLANDS' RESEARCH CENTRE T.N.O. FOR SHIPBUILDING AND NAVIGATION
Reports
No. i S The determination of the natural frequencies ofship vibrations (Dutch).
Byprof. ir H. E. Jaeger. May 1950.
No. 2 Confidential report, not published. July 1950.
No. 3 S Practicalpossibilities ofconstructional applications ofaluminium alloys to ship construction.
By prof. ir H. E. Jaeger. March I 951.
No 4 S Corrugation of bottom shell platingin ships with all-welded or partially welded bottoms (Dutch).
By prof. ir H. E. Jaeger and ir H. A. Verbeek. November 1951.
No. 5 S Standard-recommendations for measured mile and endùrance trials of sea-going ships (Dutch).
By prof. ir 7. W. Bonebakker, dr ir W. J. Muller and ir E. J. Die/il. February 1952.
No. 6 S Some tests on stayed and unstayed masts and a comparison of experimental results and calculated stresses (Dutch).
By ir A. Verduin and ir B. Burghgraef. June 1952. No. 7 M Cylinder wear in marine diesel engines (Dutch).
By ir H. Visser. December 1952.
No. '8 M Analysis and testing of lubricating oils (Dutch).
By ir R. N. M. A. Malotaux and irj. G. SmiLJuly 1953.
No. 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. Audigé. October 1952.
No. 10 5 On collecting ship service performancedata and their analysis.
By prof ir J. W. Bonebakker. January 1953.
No. 11 M The use of three-phasecurrent for auxiliary purposes (Dutch). By ir J. C. G. van Wyk. May 1953.
No. 12 M Noise and noise abatement in marine engine rooms (Dutch).
By "Technisch-Physische Dienst T.N.O.- 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. 14 M The purification of heavy fuel oil for diesel engines (Dutch).
By A. Bremer. August 1953.
No. 15 S Investigation of the stressdistribution 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. Malolaux and dr F. vañ Zeggeren. March 1957.
No. 18 M 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 irJ. C. G. van Wdk. February 1957. No. 19 M Crankcase explosions (Dutch).
-By ir J. H. Minkhorst. April 1957.
No. 20 S An analysis of the application of aluminium alloys in ships' structures.
Suggestions about the riveting betweensteel and aluminium alloy ships' structures. By prof. ir H. E. Jaeger. January l955
On stress calculations in helicoidal shells and propeller blades.
By dr ir J. W. Cohen. July 1955.
Some notes on the calculation of pitching and heaving in longitudinal waves.
By ir J. Gerriisma. December l955
Second series of stability experiments on models of lifeboats. By ir B. Burghgraef. September 1956.
Outside corrosion of and slagformation on tubesin oilfired boilers (Dutch). By dr W. J. Taat. April 1957.
Experimental determination of damping, added mass and added mass moment of inertia of a shipmodel.
By ir J. Gerritsma. October 1957.
Noise measurements and noise reduction in ships.
By ir G. J. van Os and B. van Steenbrugge. May 1957.
Initial metacentric height of small seagoing ships and the inaccuracy and unreliability of calculated curves of righting levers.
By /nof. ir J. W. Bonebakker. December 1957.
No. 28 M Influence of piston temperature on piston fouling and piston-ring wear in diesel engines using residual fuels.
By ir H. Visser. june 1959.
No. 29 M The influence of hysteresis on the value of the modulus of rigidity of steel.
By-jr A. Hoppe and ir A. M. Hens. December 1959.
No. 30 S Ari experimental analysis of shipmotions in longitudinal regular waves.
By ir 7. Gerritsma. December 1958.
No. 31 M Model tests concerning damping, coefficients and the- increase in the moments- of inertia due to entrained water
on ship's propellers.
The effect of a keel on the rolling characteristics of a ship.
By ir J. Gerrisma.Ju1y 1959.
The application of new physical methods in the examination of lubricating oils. (Continuation of report No. 17 M.)
By ir R. N. M. A. Malotaux and dr F. van Zeggeren. November 1959.
Acoustical principles in ship design.. By ir J. H. Janssen. October 1959. Shipmotions in longitudinal waves.
By ir J. Gerritsrna. February 1960.
Experimental determination of bending moments for-three models of different fûllness in regular waves. By ir J.. Ch. De Does. April 1960.
Propeller excited vibratory forces in the shaft of a single screw tanker
By dr ir J. D. van Manen and ir R. Wereldsma. June 1960.
Beamknees and other bracketed connections.
By prof. ir H. E. Jaeger and ir J. J. W. Nibbering. January 1961
Crankshaft coupled free torsional-axial vibrations of- a ship's propulsion system. By ¡r D. van Dore and N. J. Vzser, June 1963.
On the longitudinal reduction factor for the added mass of vibrating ships with rectangular cross-section. By ir W. P. A. Joosen and dr J. A. Sparenbeig. April 1961.
Stresses in flat propeller bláde models determined by the moiré-method. By. ir F. K. Ligienberg. June 1962.
Application of modem digital computers in naväl-architecture.
By ir H. J. Zunderdorp. June 1962.
Raft trials and ships' trials with some underwater paint systems.
By drs P. de Wolf and A. M. van Londen. July 1962.
Some acoustical properties of ships with respect to noise-control. Part I. By ir J. H. Janssen. August 1962.
Some acoustical properties of ships with respect to noise-control. Part II. By ir J. H. Janssen. August 1962.
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.
Results of an inquiry iñtothe condition of ships' hulls iñ relation to-fouling and corrosion.
By ir H. C. Ekama, A. M. van Landen and drs. P. de Wolf. December 1962.
Distribution of damping and added mass along the length of a shipmodel.
By prof. ir J. Gerriisma and W. Beukelman. March 1963.
The influence of a bulbous bow on the motions and the propulsion in longitudinal waves.
By prof. ir J. Gerrils,na, and W. Beukelman. April 1963.
-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 Landen and ir J. Remmells. July 1963.
The breaking of large vessels. By Prof. ir H. E. Jaeger.
A studie of ship bottom paints in particular pertaining to the behaviour and action of anting-fooling paints.
By A. M. van Landen. September 1963.
-Fatigue of ship structures.
By ir J. J. W. Nibbering. September 1963.
Communications
Report on the use of heavy fuel oil in the tanker "Auricula" -of the Anglo-Saxon Petroleum Company (Dutch).
August 1950.
-Ship speeds over the measured mile (Dutch).
By ir W. H. C. E. Rösingh. February 1951.
-On voyage logs of sea-going ships and their analysis (Dutch).
By prof. ir J. W. Bonebakker and ir J. Gerrisma. November 1952.
Analysis of model experiments, trial and service performance data of a single-screw tanker.
By prof. ir J. W. Bonebaklçer. October 1954.
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.
Approximative calculation of the effect of free surfaces on transverse stability (Dutch).
-By ir L. P. Herfst. April 1956.
-On the calculation of stresses in-a stayed mast.
By ir B. Burghgraef. August 1956.
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.
Review of the investigatiorisinto 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
