ship Collision Finite Element Analysis

D.P.Servis & M.Samuelides National Technical University of AthensShip Collision Analysis Using Finite Elements

Paper Number (x) SAFER EURORO Spring Meeting, Nantes 28 April 1999 1

SHIP COLLISION ANALYSIS USING FINITE ELEMENTS

By

Dimitris P. Servis and Manolis SamuelidesDepartment of Naval Architecture and Marine Engineering

National Technical University of AthensAthens, Greece

1. INTRODUCTION

Today's merchant ships confront a great number of operational requirements set both by the

shipping community and the international community. The requirements set by the shipping

community mainly relate to the ships' speed, building and operational cost. Additional

requirements set from both the shipping and international communities are the safety of the

ship as a mean of transportation as well as the safety of the natural environment and man-

built structures. These requirements set new methodologies for the design and operation of

ships. As sea routes are getting denser and speeds higher, there is a good possibility that a

ship may experience extreme loads during her lifetime. Further, higher speeds may cause

magnified operational loads or severe loads like slamming. Denser sea routes increase the

possibility of an accident involving ships or ships and shore or offshore structures. In

addition, as speeds along these sea routes are increased, the possibility of an accident also

increases.

There are two complementary ways in dealing with these problems. The first one is to

prevent the occurrence of extreme loads and accidents. This is being achieved using on board

sophisticated surveillance and monitoring equipment and well-trained crews. In addition the

surveillance of sea routes, especially in high traffic areas near harbours, channels and

offshore structures contribute a lot in minimising accident occurrence. The second aspect on

the solution of the problem is to mitigate the effects of a potential accident. This is done by

developing structures that may tolerate a damage within the limits of a required safety of the

structure, payload and environment. In order to achieve this, the loads induced on the

structure should be quantified.



During the last decades, the issue of impact loads on ship structures is of major concern, and

collision loads have been identified, at least in the cases of ships carrying hazardous cargo

initially and passenger ships more recently, as a phenomenon, which should be taken into

consideration in the design stage of a ship. The interest on the effect of collision loads on

passenger ships has become higher, since the advent of high speed ferries with conventional

or non-conventional hull shape. Since there is no previous experience on the behavior of

these new structures and of the new materials used in many cases, the investigation of

collision effects is imperative.

There are two major questions that naval engineers working on ship collisions should be able

to approach: One concerns the simulation of ship collisions and the prediction of the

damages, which occur during the incident. The other is the identification of collision scenario

or scenarios, which the ship under consideration should be checked against in order to assess

her capacity to withstand collision loads.

The paper addresses the former of the above mentioned questions. It is the main object of the

work reported herein, to examine the capabilities of explicit finite element codes for collision

simulation. In particular the paper reports on a) the simulation of impacts on ship structural

components, b) the simulation of small-scale collision tests and c) on the simulation of the

collision of a Ro-Ro vessel, the MS-DEXTRA.

The simulation of impacts on structural components allows investigating whether failure

modes, which have been observed during large-scale impact tests may be obtained by the

numerical analysis.

The small-scale collision tests that are simulated were performed in the University of

Glasgow in the early ‘80s. The models, which represented the parallel section of a tanker,

were either fixed or floating freely in a towing tank. The simulation of these tests could

provide necessary confidence for the numerical results.

Finally the simulation of the collision of the Ro-Ro vessel would allow examining, whether

the structural arrangement could be considered adequate to withstand the collision loads

under the defined scenario.



The FE simulation of a collision encompasses a number of individual problems, which should

be given appropriate attention. These problems are:

Ø The selection of a mesh, which should be fine enough for the results to converge and to

reproduce the failure modes and coarse enough for executing the code in acceptable time

(for the collision analysis of the Ro-Ro presented in the paper, the CPU time is counted in

days).

Ø The inclusion of the effect of the surrounding water.

Ø The extent of the ship, which would be modelled as a deformable body

Ø The modelling of the material.

Ø The material failure criterion, which appears to be a “weak point” in the collision

analysis, either this is performed with simplified analytical techniques or numerically.

The work reported herein is expected to give to the investigators some guidance in order to

assist them to approach the complicated task of the collision simulation.

2. FINITE ELEMENT CODE SELECTION AND APPLICATION TO SHIP

COLLISIONS

Currently, many researchers worldwide work on the numerical simulation of ship collisions.

Kitamoura O. et al [6], [7], Lee J.W. et al [8], and Lehmann E. et al [9] carry out outstanding

work. Very large FE element models have been developed and in some cases comparisons

with experimental results were carried out. Nevertheless, the prediction of collision results

with FE methodologies has not yet been fully proven and achieved. In the field of material

failure, Lehmann [9] has introduced a sophisticated material model, which returns quite good

results.

In the framework of selecting an efficient FE code for simulating ship collisions, several

criteria should be met. These criteria relate to the modelling capabilities for both the internal

and the external collision mechanics. In this aspect, the code must be capable of modelling

the two ship motions during and after the collision (external mechanics), as well as the

deformation and collapse of the structures (internal mechanics). In general, codes like

MSC/DYTRAN [1] and ABAQUS/Explicit, using the explicit time-integration scheme

incorporate such features. Since MSC/DYTRAN is currently available in NTUA, it was

decided to use it for the simulations. This code, licensed to the Computer Centre of NTUA is



an explicit, three-dimensional finite element code for analysing the dynamic, nonlinear

behaviour of structures. MSC/DYTRAN integrates the capabilities of DYNA3D and PISCES

codes, used in the past to simulate ship impacts. Researchers currently use MSC/DYTRAN

worldwide for modelling ship collisions. The features of MSC/DYTRAN considered

essential for solving ship collisions are discussed below:

i) External Mechanics: MSC/DYTRAN being an explicit FE code, allows for the

structures to freely move in space without defining any constraints. This feature is

essential, since it means that the structures of the colliding ships may interact and

separate as they would in a zero gravity space. Thus, no boundary conditions are

needed. In addition, there is no need to prescribe the motion of the striking ship and

the collision scenario is only defined by her initial velocity and the relative position of

the two ships in space. The program outputs the motion of both striking and struck

ships during the simulation. Another critical feature is the explicit modelling of rigid

structures and point masses. Rigid structures are used for modelling striker bows and

connections between detailed and coarse areas of the struck ship. Point masses are

used to model the displacement distribution of the actual ship and the added mass

distribution due to the movement of the struck in the water.

ii) Internal Mechanics: The available material models in MSC/DYTRAN can be used to

account for the material behaviour under dynamic loading, as much as this is possible.

Such material models take into account the material yielding and strain-rate effects on

the yield point, material hardening and failure. It is also possible to define a piecewise

linear, stress-strain characteristic. The failure criterion may be the maximum plastic

strain or defined via a user subroutine. Failure means that the element, subject to such

a deformational or stress level satisfying the failure criterion, it will lose all its

stiffness.

The above mentioned characteristics briefly state the capability of the code to model the

internal mechanics of the collision phenomenon. Nevertheless, there are certain features that

are not present and would facilitate the modelling of ship structures. These features are

mainly related to the available elements in the code. Only three- and four-node shell elements

are included, while higher degree elements could exist. Higher degree elements could be used

towards a better representation of complex hull forms and more precise representation of

stress and strain results. As much as the beam elements are concerned, the cross-section

definition of the element does not include offsets for the moments of inertia, importing a



small error in the determination of the stiffness. In addition, only special cross-section beam

elements may incorporate the effect of strain-rate hardening. This excludes the general-

purpose beam elements from the modelling of ship structures, since the strain-rate effect is of

a great importance in ship collisions. It should be noted though, that based on the available

information neither ABAQUS/Explicit supports some of these features. It is likely that these

features are not present in favour of a minimal solution time. Explicit codes calculate the time

step for every cycle based on the length of the smallest element, the sound speed in the

material assigned to this element and a scale factor. This produces a very small time step and

consequently long solution times. Higher degree elements with more nodes and more

complex formulation would cause a considerable increase in solution time.

In order to evaluate MSC/DYTRAN's ability to model ship collisions, a simplified model is

developed and tested in different structural configurations. Structural description and results

can be found in [2, 3, 17]. This model consists of a simple V-shaped striker bow and a hollow

side shell structure of the struck ship. The striker bow is rigid and its mass is equal to the

mass of the whole ship (see Figure 2.1). Point masses are used to model the actual mass of

the ship and the added mass. The material model is elastic-perfectly plastic with a maximum

plastic strain failure criterion. The various structural configurations differ in the way of

modelling the remaining parts of the struck ship and the added mass.

Figure 2.1: Simplified model for ship-ship collision

This model was used to investigate and evaluate the capabilities of MSC/DYTRAN in

modelling collisions. Comparing the results for each configuration, they exhibit a consistency

in expected vs. obtained results. As a concluding remark on these benchmarks,

MSC/DYTRAN may be considered suitable for modelling ship collisions.



3. SIMULATION OF IMPACTS ON SHIP STRUCTURAL COMPONENTS

The simulation of impacts on structural components is considered essential in many ways.

Structural components may be used in to investigate the effect of an impact in terms of

structural configuration and material used. They can also be used to evaluate modelling

techniques, such as elements used, material properties and mesh densities.

In the way of simulating the structural response of deck and side shell during a collision, two

FE models of components have been developed. The purpose of these models is to achieve

certain collapse modes observed in full-scale collision tests and investigate the effect of

different mass and collision speed of the striking structure. These models include the collision

of a dihedral bow with deck and sideshell and the collision of a hemisphere bulb against a

panel.

The first model presented here consists of a dihedral bow striking part of a sideshell and deck

(see Figure 3.1). The panel consists of large web frames and longitudinal stiffeners. The

material is elastic-perfectly plastic steel and the thickness of all components is 10mm.

Figure 3.1: The dihedral bow and the sideshell-deck panel collision model.

The panel is clamped on its perimeter - except the contact side - while the bow is rigid and

free to move in space. The translational velocity of the centre of gravity of the bow is 2.57

m/sec along the z-axis. Preliminary runs have been performed, in order to check the mesh's

efficiency. In these runs, every element of the initial mesh has been split into four elements.



In that way, the original number of elements is quadrupled. As can be seen in figure 3.2 the

results for both meshes are quite alike.

0.00E+00

5.00E-02

1.00E-01

1.50E-01

2.00E-01

2.50E-01

0.00 0.02 0.04 0.06 0.08 0.10

Time [s]

Dis

plac

emen

t [m

]

meshmesh4

0.00E+002.00E+044.00E+046.00E+048.00E+041.00E+05

1.20E+051.40E+051.60E+05

0.00 0.02 0.04 0.06 0.08 0.10

Time [s]

Def

orm

atio

n E

nerg

y [J

]

meshmesh4

Figure 3.2: Deformation Energy vs. Time and Bow Displacement vs. Time for different mesh

densities

Mesh evaluation is a procedure that has proved critical for the collision modelling. This field

is still under investigation and considerable conclusions on how collisions should be

modelled have been reached using such techniques.

The runs performed for the above-mentioned model are set up using different bow masses.

The initial mass of the bow is equal to the total mass of the deformable structure. Runs are

performed for 1, 2, 3, 4, 6, 8 and 10 times the initial bow mass. For these runs, the

elastoplastic material's yield stress is independent to the strain rate. Additional runs are

performed for 10 times the initial bow mass, but with strain-rate dependency according to the

Cowper-Symonds rule. For these runs, the deformation energy absorbed by the deck and

sideshell structure is plotted against the displacement of the bow (figure 3.3). It is assumed

that the displacement of the bow represents the deformation of the structure. The bow is rigid

and thus non-deformable. In addition it can only move towards the structure or away from it

(that is, along the z-axis). It cannot rotate or move otherwise.



0 . 0 0 E + 0 0

1 . 0 0 E + 0 5

2 . 0 0 E + 0 5

3 . 0 0 E + 0 5

4 . 0 0 E + 0 5

5 . 0 0 E + 0 5

6 . 0 0 E + 0 5

7 . 0 0 E + 0 5

0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8

D i s p l a c e m e n t [ m ]

Def

orm

atio

n E

nerg

y [J

]

W

2 X W

3 X W

4 X W

6 X W

8 X W

1 0 X W

1 0 X W S R

Figure 3.3: Displacement vs. Deformation Energy

As can be seen in figure 1.2.2.3, in all cases except the strain-rate dependent case (10XW

SR), the displacements vs. deformation energy curves coincide. Additional runs have been

performed at this point in order to determine whether the velocity affects the solution.

Towards that direction, for the 10 times the initial bow mass case, the initial velocity of the

bow was increased and its mass was set to the initial value (1XW) in such a way that the

initial kinetic energy of the bow would remain the same. In figure 3.4, displacement vs.

deformation energy is plotted for four cases: large mass and large velocity, with or without

taking into account the strain-rate effect.

0.00E+00

2.00E+05

4.00E+05

6.00E+05

8.00E+05

1.00E+06

1.20E+06

0.00 0.10 0.20 0.30 0.40 0.50 0.60 0.70 0.80 0.90 1.00

Displacement [m]

Def

orm

atio

n E

nerg

y [J

]

Large mass 10X

Large mass 10X SR

Large Velocity 10X

Large Velocity 10X SR

Figure 3.4: Displacement vs. Deformation Energy



Figure 3.4 shows that in the case where no strain-rate dependency is present, the results are

relatively closer. On the contrary, if strain rate is taken into account, the deformation energy

is greater in the case of greater initial velocity, for the same displacement.

The second model investigated simulates the collision of a hemisphere bulb against a panel.

In this case, the panel mentioned previously is subject to a collision with a hemisphere bulb

(figures 3.5 and 3.6). In this case, the longitudinal stiffeners and sideshell have been

removed. The diameter of the bulb is a little greater than the web frame spacing. The purpose

of this model is to examine the ability of the bulb to pierce the plate, observe the results on

the web frames and determine the conditions under which tripping of the web frames may

occur. The material of the panel is considered elastic-perfectly plastic and strain-rate

dependency of the yield stress is considered.

Figure 3.5: Hemisphere bulb in collision with panel.

Figure 3.6: Hemisphere bulb in collision with panel.

The panel is simply supported along the edges parallel to the y-axis.



The runs performed with this model can be divided in two groups: one group where the bulb

strikes the panel exactly on its geometric centre of gravity, and one where it is shifted

towards the negative x-axis. For both groups, there are two initial velocities considered for

the bulb: 7m/s and 3.5 m/s.

Figure 3.7: Strain results in time 0.5s for 7m/s, central and shifted strike

In the above figure, it can be seen that in the case where the bulb strikes on the centre of

gravity of the panel, there is no considerable bending along an axis parallel to the web

frames. On the contrary, in the case where the bulb is shifted, bending occurs in both

transverse directions of the plate. It can be also seen that in the case of the central strike, the

elements of the contact region fail and the bulb pierces the plate, whereas in the case of the

non-central the elements near the web frame fail and a flap-like part of the contact region is

cut off. This behaviour has also be observed in large-scale collision tests performed by TNO

(see fig. 3.8).

Figure 3.8: Results on large-scale plate with web frames.



4. SIMULATION OF COLLISION EXPERIMENTS

In order to validate the collision simulation capabilities and verify the extensionality of the

results, collision experiments [4] were simulated using FE. It has been observed that material

behaviour is an issue of great importance in the case of ship collisions. Material property

definition is at this time under investigation. As well as the material definition, a lot of work

is done on establishing mesh evaluation techniques and result consistency checks. The FE

model used to perform these tasks is a box-like structure with longitudinal and transverse

bulkheads (figures 4.1 to 4.4).

Figures 4.1 and 4.2 present the experimental set-up and struck ship's main dimensions. The

striking bow is a sharp rigid dihedral. The experimental set-up allows for the initial velocity

of the bow prior to the collision to be calculated. Two sets of experiments were performed:

one where the struck ship is clamped along its ends and one where the struck ship is free to

float in water. Tests are performed for different bow masses and velocities and struck ship

masses. The initial condition for this run is the striking vessel's initial velocity, equal to

3.4m/s, while it's mass is 28.6kg.

4.1 SIMULATIONS PERFORMED WITH THE STRUCK VESSEL CLAMPED

Two FE models are developed for this case. One for the whole struck vessel and one for the

area bound by two consecutive transverse bulkheads and a longitudinal bulkhead. Different

material definitions and mesh densities have been used for the several runs performed. These

materials include a linear elastic material, an elastic-perfectly plastic material, an elastoplastic

material and an elastoplastic with strain-rate dependency material. In Table 4.1 a result

summary for different materials and mesh densities is presented, along with experimental

results. The results are considered quite satisfactory. The more the material model approaches

a realistic behaviour, the more the results converge to the experimental values. For the full

model, the permanent deflection appears to be the same with the experimental result.

Deformational energy is also quite close, considering that the experimental result is the

subtraction of the final kinetic energy of the striker from the initial kinetic energy. In

addition, good convergence is achieved using different mesh densities. This is also obvious in

Figures 4.5 and 4.6. In the former, code names used to describe each curve correspond to

numbers 1-8 in Table 4.1. It can be seen that the curves for different mesh densities almost

coincide.



3800.00

1692

.00

30°

Figure 4.1: Experimental set-up.

4in

1'

1'

4'

R 2in

Figure 4.2: Struck model main dimensions



Figure 4.3: Struck model: coarse and dense mesh

Figure 4.4: FE model of the whole structure

M.Samuelides & D.P.Servis


Experiment Elastic (1)* Elastic-Perfectly Plastic Elastoplastic

Coarse

Mesh(2)

Dense Mesh(3) Coarse Mesh

(4)

Dense Mesh (5)

Collision Duration [s] 0.0106 0.0132 0.0133 0.0123 0.0126

Maximum Force [N] /

Time Occ.[s]

47150.85/

0.005

13495.5/

0.0098

13158.49/

0.0101

16888.82/

0.0096

16814.88/

0.097

Final Deformational

Energy [J]153 0.520 158.166 158.562 155.407 156.036

Percentage of the

initial Energy[%]93% 0.31% 95.68% 96.01% 94.01% 94.57%

Deviation from

Experiments[%]-99.66% 3.38% 3.64% 1.57% 1.98%

Permanent

Deflection [m]0.0134 0.0003 0.0205 0.0222 0.0192 0.0204

Deviation from

Experiments[%]52.98% 65.67% 43.28% 50.7%

Final striker's

Velocity [m/s]-0.9 -3.391 -0.644 -0.616 -0.802 -0.783

Deviation from

Experiments[%]-28.41% -31.6% -10.84% -13%

Table 4.1: Result summary

*Although the elastic material inappropriate for the simulation of the tests, it is included in the evaluation process to examine the energy equilibrium during collision, the

behaviour of the contact area and the determination of the deformation energy. The results for this case are presented in Table 1.2.2.1 just for completeness, though it would

be nice if ships could have such a response in actual collisions.



Force vs. Displacement

0.00E+00

5.00E+03

1.00E+04

1.50E+04

2.00E+04

2.50E+04

3.00E+04

3.50E+04

4.00E+04

4.50E+04

5.00E+04

0.000 0.005 0.010 0.015 0.020 0.025

Displacement [m]

Forc

e [N

]

sideresel (1)sideresey (2)

sidereseyX2 (3)sidereseyX3

sidereseyh (4)sidereseyhX2 (5)

sidereseyhsr (6)sidereseyhsrX2 (7)

dry reseyhsr (8)

Figure 4.5: Force vs. Bow Displacement curve

Force vs Bow Displacement

0.00E+00

2.00E+03

4.00E+03

6.00E+03

8.00E+03

1.00E+04

1.20E+04

1.40E+04

1.60E+04

0.00 0.01 0.01 0.02 0.02 0.03

Bow Disp. [m]

Forc

e [N

]

sideresey

sidereseyX2

sidereseyX3

Figure 4.6: Force vs. Bow Displacement curves example for different mesh densities

Force vs. bow displacement or indentation curves can be used as a convergence criterion

for the mesh density. This is extremely convenient, since the small solution time allows

for the impact simulation of several structural components of the struck area using

different meshes. In this way, convergence may be achieved and the appropriate mesh

density can be selected.



Figure 4.7: Von Mises stress distribution

The whole model impact simulation is also used for the estimation of the critical

modelling dimensions for a collision model. As can be seen in figure 4.7, significant Von

Mises stresses extend along three compartments of the struck ship. The side shell and

internal structure do not plastify in these regions but considerable membrane tension in

the side shell and bending of the transverse bulkheads adjacent to the impact area is

observed.



4.2 SIMULATION WITH THE STRUCK SHIP FREELY FLOATING

In this second set of experiments, the struck ship is free to float on the water surface. The

experimental setup does not differ a lot from the previous one. In table 4.2 the initial

conditions for both the experiments and corresponding simulations are defined.

Case Striking mass

[kg]

Struck mass

[kg]

Striking

velocity [m/s]

1 55.4 39.0 2.62

2 55.4 39.0 2.89

3 55.4 39.0 2.89

4 55.4 39.0 3.1

Table 4.2: Initial conditions for the tests

Additional mass was included in the form of node concentrated masses on the

longitudinal bulkheads. These masses simulate the additional weight due to measuring

devices. Such concentrated masses were also used to simulate the hydrodynamic added

mass. This mass is allocated on the side shell and boundary bulkheads up to the draught

of the floating ship in still water. A value of 40% of the struck mass was accounted for

this purpose, which is the usual case, but was also verified for the experimental model. In

tables 4.3 and 4.4 some experimental and numerical results are presented.

1

3

4

2



Experimental Numerical

Test Sway

velocity

[m/sec]

Yaw

velocity [s-1]

Deflection

[mm]

Sway

velocity

[m/sec]

Yaw

velocity

[s-1]

Deflection

[mm]

1 1.90 0.00 8.3 2.246 0.000 7.1

2 1.55 2.20 8.1 2.085 3.020 6.5

3 Not acquired due toinstrument failure

8.4 2.420 0.000 8.0

4 1.75 3.00 11.4 2.195 3.120 7.8

Table 4.3: kinematic results for the experiments and numerical simulations

Test Struck

kinetic

energy [J]

Striker

kinetic

energy [J]

Deformation

energy [J]

Collision

duration

[s]

Maximum

force [N]

Time of

occ. [s]

1 89.24 4.91 44.78 0.015 16704.26 0.0078

2 102.33 20.22 50.36 0.014 17140.55 0.0075

3 103.53 7.47 61.09 0.014 18331.94 0.0084

4 114.01 25.10 62.12 0.013 18165.36 0.0073

Table 4.4: Energy results for the numerical simulations

It can be seen in table 4.3 that the final velocities calculated by the numerical models are

greater than the experimental ones. This can be due to the approximation of the effect of

the surrounding water with concentrated masses. A concrete answer to this may be given

by combined fluid-strucutre and structure-structure interaction simulations (CFSS)

already in progress. The permanent deflection at the collision area is greater in all

experimental cases than the numerical ones. The difference though is not very significant

in the cases where a central impact is performed. This can also be due to the modelling of

added mass.



Case Initial

Energy

[J]

Maximum

Deformational

Energy [J]

Plastic

Deformational

Energy [J]

Elastic

Deformational

Energy [J]

Dry 165.308 164.238 115.818 48.420

Wet 165.308 94.075 51.173 42.903

Wet 1 190.144 91.601 44.782 46.819

Wet 2 231.353 93.924 50.364 43.559

Wet 3 231.353 111.197 61.086 50.112

Wet 4 266.197 107.884 62.121 45.764

Table 4.5: Energy results for the numerical simulations

In table 4.5 energy results from numerical simulations are presented. The two first cases

correspond to the whole model impact simulation when it is clamped and to an equivalent

simulation with the boundary conditions removed. It can be seen that the elastic

deformational energy is almost the same for all cases though the plastic deformational

energy varies significantly. It is also observed that the amount of elastic energy absorbed

in the cases where the struck ship is free to float is significant compared to the plastic

deformational energy. The elastic energy is absorbed by the side shell and bulkheads,

especially in their intersections (figure 4.8).

Figure 4.8: Von Mises stress distribution



This observation indicates that the full breadth of the test model should be modelled in

order to include all its significantly loaded areas. Combining this observation with the

previous conclusion on the modelling length, a collision FE model should at least extend

along three compartments and the full breadth of the ship. The areas apart from the struck

region absorb elastic deformational energy. If mostly the sideshell or bulkheads will

absorb this energy depends on the inertia forces exerted on the ship due to its movement.

In the case of a complicated ship structure, in order to reduce the number of elements and

nodes, equivalent formulations to the stiffened plates, such as orthotropic plates or

aggregated stiffeners, could be used for areas away from the collision area. A suitable

formulation based on orthotropic plates theory is under investigation.

5. SIMULATION OF COLLISION OF A LARGE RO-RO

Within DEXTREMEL - Design for Structural Safety under EXTREME Loads - project

the MS-DEXTRA has been modelled with finite elements. All the information and

drawings used for the development of the FE model were provided by ASTILLEROS

ESPAÑOLES [2], [5] as part of the deliverables of the project DEXTREMEL The

intention of this model is to use it in global collision runs. The analysis being done

includes the assessment of its structural integrity after the collision, definition of

optimum modelling of the panels at the impact region and possible substitution of other

panels with equivalent orthotropic plates. In this model, only a central part of the ship is

modelled in detail. Since the target vessel is a passenger ferry, no actual parallel body

exists. Nevertheless, at this stage it was selected to model part of the ship as it was

parallel, according to the midship section. The modelling length is about 48 structural

frames, beginning from frame 66 (bulkhead) until frame 114 (bulkhead) with the LCG

being at about frame 97.5. The corresponding length in the bodyplan begins at about

station 12.5 and ending at station 16.5. The bulkhead positions are at frames 66, 89, 102

and 114. The hull module modelled includes three compartments: the engine room and

part of the machinery room. The whole breadth of the ship is modelled in detail.



Figure 5.1: The Target Vessel FE model

Figure 5.2: Web Frame and Ordinary Frame



Figure 5.3: Bulkhead and Bow

In the previous Figures, views and sections of the target vessel are presented. In Figure

5.3 a FE of a bow is also presented. This is the bow of the target vessel and is used as a

striking bow. The target vessel FE model consists of 36391 nodes and 65811 shell and

beam elements. Shell elements were used for the sideshell, bulkheads, web frames and

stringers. Beam elements were used for all longitudinal stiffeners and upper decks'

stiffeners (figures 5.1 to 5.3). The struck ship for and aft of the three compartments,

which were modelled in detail, is modelled with elastic beams of the same area and

moments of inertia as the midship section. These beams are connected to the hull module

at the centre of gravity of the midship section using a rigid frame. In that way, the elastic

beams simulate the inertia effect of the rest of the ship on the hull module, as forces are

applied on it through the rigid frames. It was selected to use this configuration since the

standard rigid hull does not allow for horizontal bending of the ship and rotation of the

boundary bulkheads. The mass moments of inertia of the rigid bow are selected so that it

can only move along the x-axis (figure 5.3). Both ships are floating at the same draught,

but the struck vessel is in light ship condition. At this stage no added hydrodynamic mass

is used. The material used for the struck model is mild steel. The material model is

elastoplastic and includes material hardening but no strain-rate hardening. The failure

criterion is the maximum plastic strain. When the mean plastic strain of the element

reaches the value 0.18 then the element fails.



The results presented are acquired from a preliminary run. The following figures show

some equivalent stress results. These were recovered at times 0.15 sec, when the impact

force has a local maximum and 0.20 sec when there is extensive rupture of the side shell

due to bulb piercing. The images are rendered and shaded in order to make different

failure modes recognisable.

Figure 5.4: Relative position of the struck area and striking bow at 0.15secs: Equivalent

stress

Figure 5.5: Overall damage of the struck area at 0.15secs: Equivalent stress



Figure 5.6: Inside view of the damage at 0.15secs: Equivalent stress

Figure 5.7: Inside view of the damage caused by the bulb at 0.15secs: Equivalent stress

Tripping ofthe stringer

Tripping ofweb frame

Tripping ofOrdinaryframe



Figure 5.8: Views of the damage at 0.20secs: Equivalent stress



Figure 5.9: View of the damage caused by the bulb at 0.20secs: Equivalent stress

A first remark is that the selection of a rigid striking bow, especially of the kind used for

this analysis, may give an exessive estimation of the damage. The selected bow has a

bulb and a significant flare. In that way the contact force is not evenly distributed along

the contact area – as it would with a nearly vertical bow. Of course, this is an actual bow

shape that the ship may encounter in a collision accident dyring her lifetime. The shape of

the bow and its rigidity cause very localised rupture. Further, at the upper part of the

contact area (figures 5.6 and 5.8) the rigid and sharp bulwark causes excessive horizontal

tearing of the sideshell; a damage that would had been different in case of a deformable

striking bow.

Before the rupture of the sideshell at the lower struck area, typical failure mechanisms

can be identified. Tripping of frames and stringer occurs as well as crushing of the web

frames near the main deck (figure 5.7). In figure 5.8 it can be seen that the extent of the

modelled structure seems to be sufficient, in order to obtain the stress field 0.2 sec after

contact. The stresses in areas apart from the struck one are below the first yield point and

seem to be greater along the intersection of the deck with the side shell (figure 5.8).



6. CONCLUSIONS

The paper refers to the simulation of ship collisions. The following related aspects are

considered to be essential for such a task:

(i) Selection of mesh: The selection of the mesh has been based on the convergence of

the force – indentation curve, and on its adequacy to reproduce failure modes

observed in impact tests.

(ii) Effect of surrounding water: The surrounding water is modelled using concentrated

mass elements distributed over the wetted area, if a detailed model of the hull is

available. From the comparison of experimental and numerical results, it seems that

this formulation is adequate if the impact longitudinal position is close to the LCG

of the ship.

(iii) Extent of the detail model of the impacted hull: The modelling extent of the struck

ship should cover the plastic contact area and the elastically stressed areas in the

compartments, which are adjacent the heavily deformed contact area. Further there

are indications that an elastic stress field may be present over the whole breadth of

the model, and therefore it is considered appropriate to model the whole structure in

between two cross-sections. Equivalent orthotropic panel formulation should be

investigated in order to determine whether such panels can be used in areas away

from the struck region.

(iv) Failure criterion: The failure criterion, which has been used is based on the plastic

strain of an element. The value of the maximum plastic strain, which has been

selected is 18%. Such a value seems to result in realistic failure modes and in

results, which are close to experimentally obtained values of penetration. More

effort is needed in order to select a failure criterion, which would be used with

more confidence in the collision simulation.

The results obtained from the finite element simulation, will be used a) for the assesemnt

of the of collision behaviour of a ship under the defined collision scenario, b) for the

relative comparison of structural arrangments, and c) for the validation of analytical



techniques for collision analysis. Within the DEXTREMEL project the analytical

technique will be developed by DTU [18].

7. ACKNOWLEDGMENT

The work is part of the DEXTEMEL, Brite-Euram project financed by the Directorate –

General XII of the European Commission. The Commission is thanked for its support.

8. REFERENCE

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ship Collision Finite Element Analysis