Link - Sobena

Transcrição

Link - Sobena

S O B E N A
Marine Systems & Ocean Technology
Journal of SOBENA
www.sobena.org.br/msot
Adress: Av. Presidente Vargas, 542 - Grupo 709 a 713 - Centro - Rio de Janeiro - RJ - Brasil - CEP 20071-000
Telephones: [+55](21) 2283-2482 - Telefax: [+55] (21) 2263-9079 - E-mail: [email protected] - Site: www.sobena.org.br
List of Editors
Júlio Romano Meneghini
Universidade de São Paulo, Brazil
[email protected]
Celso Pupo Pesce
Universidade de São Paulo, Brazil (Chief-Editor)
[email protected]
Torgeir Moan
Norwegian University of Science and Technology, Norway
[email protected]
Marcelo de Almeida Santos Neves
Universidade Federal do Rio de Janeiro, Brazil (Chief-Editor)
[email protected]
Michael M. Bernitsas
University of Michigan, USA
[email protected]
Belmiro Mendes de Castro Filho
[email protected]
Günther Clauss
Technical University of Berlin, Germany
[email protected]
Paulo de Tarso Temístocles Esperança
Universidade Federal do Rio de Janeiro, Brazil
[email protected]
Segen Farid Estefen
[email protected]
Helio Mitio Morishita
[email protected]
Celso Kazuyuki Morooka
Universidade de Campinas, Brazil
[email protected]
Kazuo Nishimoto
[email protected]
Apostolos Papanikolaou
National Technical University of Athens, Greece
[email protected]
Floriano Carlos Martins Pires Jr
[email protected]
Claudio Ruggieri
[email protected]
Odd Faltinsen
Norwegian University of Science and Technology, Norway
[email protected]
Claudio Mueller Prado Sampaio
[email protected]
Jeffrey M. Falzarano
Texas A&M University, USA
[email protected]
Alexandre Nicolaos Simos
[email protected]
Antonio Carlos Fernandes
[email protected]
Sergio Hamilton Sphaier
[email protected]
José Alfredo Ferrari Jr
Petrobras, Brazil
[email protected]
Célio Taniguchi
[email protected]
André Luiz C. Fujarra
[email protected]
Eduardo A. Tannuri
[email protected]
Carlos Guedes Soares
Universidade Técnica de Lisboa, Portugal
[email protected]
Pandeli Temarel
University of Southampton
[email protected]
Atilla Incecik
Universities of Glasgow & Strathclyde, UK
[email protected]
Armin Walter Troesch
University of Michigan, USA
[email protected]
Breno Pinheiro Jacob
[email protected]
José Márcio do Amaral Vasconcellos
[email protected]
Jan Otto de Kat
A. P. Moeller-Maersk, Denmark
[email protected]
Dracos Vassalos
University of Strathclyde, United Kingdon
[email protected]
Carlos Antonio Levi da Conceição
[email protected]
Murilo Augusto Vaz
[email protected]
Clóvis de Arruda Martins
[email protected]
Ronald W. Yeung
University of California at Berkeley, USA
[email protected]
Volume 5
Number 2
December 2010
Chief-Editors
Marcelo de Almeida Santos Neves
Universidade Federal do Rio de Janeiro
Celso Pupo Pesce
Universidade de São Paulo
JOURNAL OF
SOBENA
Sociedade Brasileira de Engenharia Naval
A
Aims and Scope
The design process of marine systems is one of formulation, evaluation and modification. Very often the problems confronting the
designer are effectively complex problems, particularly on the technical side. Analytical models have to be invoked and applied together
with numerical and experimental simulations, guided by intelligent experience, at all levels of the design chain.
In the past these difficulties have been more concentrated on few particular types of marine vehicles and systems. In particular,
naval architects have designed surface ships. Specialised methodologies and rules have been developed and accumulated in this
field. Some excellent periodicals are dedicated to the coverage of researches and developments in this sector.
More recent technological developments, particularly in the offshore industry, have challenged this knowledge, introducing many,
and often radically distinct departures from the more conventional designs. Hence, largely multidisciplinary technologies are presently at the frontline, demanding fresh contributions not only from the naval architecture and ocean engineering fields, but also
from all contributing areas as civil, mechanical, electrical, material, petroleum, coastal and oceanographic engineering, applied
oceanography and meteorology and applied mathematics.
Marine Systems & Ocean Technology intends to contribute to this wide and rich technological scenario by providing a forum for
the discussion of mathematical, scientific and technological topics related to:
• hydrodynamic and structural analysis of any fixed and floating marine systems (including ships and advanced marine vehicles),
• underwater technology (including submarines, robotics, design and operation of diving systems, surveys and maintenance
systems, umbilical cables, pipelines and risers),
• computational methods in naval architecture, offshore/ocean engineering, coastal engineering and related areas,
• environmental studies associated with oil spills and leakage prevention and control, safety concepts and risk analysis applied
to marine systems, wave-energy extracting devices and sea resources in general,
• ocean and river transportation economics, marine engineering and environmental protection, offshore support bases, offshore
logistics.
Marine Systems & Ocean Technology is an editorial initiative jointly coordinated by SOBENA and CEENO. SOBENA is an abreviation for Sociedade Brasileira de Engenharia Naval, a learned society founded in 1962 for promoting technological development.
CEENO is a Scientific Network on Naval Architecture and Ocean Engineering organized in 1999 by leading members of the Brazilian
scientific community afiliated to two universities and two research centers: COPPE/UFRJ, USP, IPT, CENPES.
Marine Systems & Ocean Technology (ISSN 1679-3962) is published twice a year and is owned by Sociedade Brasileira de Engenharia
Naval - SOBENA, and is distributed freely to members. Rate for 2011 is R$ 200.00 for institutions and R$ 100.00 for individuals.
Issues are airmail shipped. All subscriptions are payable in advance and entered on an annual basis.
Copyright © 2005 by Sociedade Brasileira de Engenharia Naval. Printed in Brazil. Authorization to photocopy articles may be
granted by Sociedade Brasileira de Engenharia Naval, provided the material is used on a personal basis only. The Society does not
consent copying for general distribution, promotion, for creating a new work or for resale. Permission to photocopy articles must
be requested to the SOBENA main office.
A
Simulation-based design for efficiency, safety and comfort
Karsten Fach and Volker Bertram
FutureShip - A GL Company, [email protected]
Abstract
The paper surveys the role of modern computer simulations to support the design of ships. At the centre of most efforts is the quest
for fuel efficiency and today also reduction of emissions. However, designs need also to consider safety and for some ships comfort
aspects. The employed simulation techniques are frequently the same and several examples will show how simulations are used
to simultaneously assess more than one of the above design aspects. Simulation areas covered include resistance and propulsion,
seakeeping, aerodynamic flows and structural analyses. For resistance and propulsion, cavitating flow analyses with transient CFD
analyses are taken as an example, where safety aspects (structural erosion of rudders) and efficiency aspects are covered in one
simulation. CFD applications to appendages illustrate how a common simulation approach may be used to make ships more fuel
efficient or address comfort issues in vortex induced vibrations. In aerodynamics, two applications of the same code are shown:
one addresses the flow field on a helicopter deck for safe operation, the other a formal optimisation of a funnel design to minimize
smoke dispersion on a megayacht. For ships, modern computer applications help minimizing structural weight while still maintaining
safety standards as prescribed by Class Rules. The diverse applications can be reduced to three design goals (basically one design
goal with another goal acting as constraint) and a few multi-purpose simulation tools. Recent case studies taken from GL Group
experience illustrate the various applications and the common denominators.
Keywords
Simulation, ship design, CFD, FEA
1
Introduction
Ship design is increasingly supported by such simulations. Traditionally, ship design has been based on experience. This is still true
to some extent, but increasingly we rely on “virtual experience” from dedicated and well chosen simulations. Scope and depth of
these simulations guiding our decisions in design and operation of ships have developed very dynamically over the past decade.
Commercial ship design is generally “ship design for efficiency and economy”, Schneekluth and Bertram (1998). The economy
of passenger vessels is closely linked to passenger comfort, not only (fuel) efficiency. And safety aspects pose constraints for ship
designs. For example, lighter ship structures reduce steel cost in construction and fuel cost in operation, but safety considerations
require certain thickness of structures. Simulations aid here to find highly efficient solutions without compromising safety, but giving
detailed insight and allowing thus reduced margins.
The word simulation is derived from the Latin word “simulare” which can be translated as “to mimic”. The Oxford dictionary defines
“to simulate” as “to imitate conditions of a situation or process”, specifically “to produce a computer model of a process”. In this
sense virtually all computer models used in the design of ships would qualify as simulations.
Simulations allow:
• performing extended “what if” analysis on all system variables guiding design and operation
• understanding where and why problems occur
• evaluating the impact of potential investments on total system costs and performance
• supplying accepted engineering proof for alternative designs
• ...
“Classical” applications of computer simulations for ships are CFD (computational fluid dynamics) and FEA (finite-element analyses).
Both have been used for several decades to support ship design, but today’s applications are far more sophisticated than 20 years
ago. The following will review different simulation fields as found in the work of Germanischer Lloyd, showing how advanced engineering simulations have drifted from predominantly safety related applications to applications concerning comfort and efficiency.
Submitted to MS&OT on May 05 2010. Accepted on Sep 08 2010. Editor: Marcelo A. S. Neves.
Vol. 5 No. 2 pp. 61-66 December 2010
61
2
Problem description
2.1
Resistance & propulsion
Ship propulsion accounts typically for the major part of the fuel
consumption of a ship. For a large containership, for example,
90% of the fuel consumption is due to ship propulsion and
only 10% for the onboard consumers. The power required for
propulsion is determined by the resistance of the ship and the
efficiency of the propeller. We refer to Bertram et al. (2009)
for a detailed discussion of how the individual resistance components can be improved and assorted losses in propulsion can
be reduced. We focus here only on some particular interesting
applications of modern simulation technology.
In the 1990, CFD (computational fluid dynamics) was introduced to ship design became a widely accepted and used tool in
the following decade, Bertram (2000). The standard procedure
for modern hull design investigated numerically up to 10 form
variants, often successively after studying flow details. Progress
in computer power and software allows today formal optimization of hulls and it is expected that these will become state
of the art within the next decade. Hull lines can be formally
optimized for fuel efficiency or other criteria, Fig.1, e.g. Abt
and Harries (2007), Oossanen et al. (2009). Here, parametric
hull description, free-surface flow simulations and formal
optimization are combined with massively parallel computer
architectures (more than 500 processors in our case) to improve
hull shapes in short turn-around times.
A similar approach can be used to optimize the trim of a built
ship. Here, essentially the underwater hull shape is modified
by changing trim rather than hull lines. For each draft and
speed, there is a fuel-optimum trim. Trim optimization has
been proven to result in considerable fuel savings (typically
5% as compared to even keel) for relatively low investment,
Fig.2, Hansen and Freund (2010).
Fig. 1
62
Hull line optimization of offshore supply vessel. Efficiency gain
was 15% in this extreme case; 4-5% efficiency gains are typical
values for most cargo ships
Fig. 2
User interface for “ECO trim assistant” advising on optimum trim
for fuel efficiency; 2.5% savings in this case
CFD based on viscous flow models (RANSE codes) is the
most appropriate tool to support practical design of aftbody
and all appendages, Fig.3. Optimization of the aftbody lines
and appendages (e.g. wake equalizing nozzles) requires
considerably higher computer resources due to the dominant effects of viscosity and turbulence. However, pilot
applications show the feasibility of the approach and formal
optimization of aftbody lines is expected to appear soon as
a standard option in ship design.
For appendages with high local flow velocities, cavitation
may occur. Cavitation reduces efficiency, but may also pose
threats for structural integrity. Initially, reports of eroded
rudders on large containerships motivated Germanischer
Lloyd to extensive numerical studies on flows involving
cavitation. The extensive experience gained on these safetyrelated simulations was later naturally transferred to fuel
efficiency studies, for example for low-cavitation fuelefficient rudder designs, Fig.4.
Fig. 3
CFD for wake prediction, capturing hull-propeller-rudder
interaction at full scale
cabin of a megayacht. Based on the measured frequency
of the vibrations, propeller and engines were ruled out as
potential suspects. This left vortex induced vibrations. CFD
simulations of the viscous flow around the ship with all
appendages revealed locations of vortex shedding and the
computed time-histories of pressures on the appendages gave
the associated frequencies, Fig.5. This allowed pinpointing
rapidly the source of the vibration problem. (In this case the
outer propeller tunnel). Then the problematic appendage can
be redesigned and its new vortex pattern can be analysed
again. The pressure fluctuations are used as input in FEA
(finite element analyses) vibration analyses, quantifying
vibration amplitudes in the ship structure, which were
negligibly small after the design modification.
2.2
Seakeeping computations may employ various approaches, differing in computational expense and how accurately they capture
different physical phenomena, Bertram (2000). Traditionally,
Germanischer Lloyd has been very active in developing and
applying seakeeping codes, as accurate load prediction forms an
integral part of accurate structural analyses. For unconventional
designs, direct simulation is the only way to determine the loads.
The lightweight design of the record-breaking trimaran Earthrace
was only possible using sophisticated CFD simulations for the
loads and subsequent finite-element analyses for the composite
hull structure, Ziegler et al. (2006). Advanced simulations also
play a role for large glass architectures on modern cruise ships.
The adhesive bonding for the glass panels needs to be designed
strong enough for the safety (and of course comfort) of passengers, Fig.6. The same simulation technology can be applied to
predict the attainable speed in waves, or the added resistance in
waves for a given speed. Recently, this was applied to a solar
powered catamaran, Fig.7. Here the added power consumption
due to waves was important to know already in the design stage.
Seakeeping affects safety, comfort and efficiency and often more
than one item must be considered. Heave accelerations have a
major impact in seasickness (and thus indirectly attainable speed
or revenue due to passengers staying away), roll accelerations
on the integrity of lashing systems, motions on added resistance
and thus fuel consumption. Again, simulations can be combined
with optimisation, for example for optimum weather routing, e.g.
Rathje and Beiersdorf (2005).
Fig. 4
CFD for low-cavitation, high-efficiency rudder
Fig. 5
Appendages on aftbody of fast monohull (left) with associated
vortex generation (right)
Misalignment of brackets and other appendages results
in vortex shedding and associated increased resistance.
The periodic vortex shedding may also induce vibrations
(VIV = vortex induced vibrations) which pose problems
for structural integrity (fatigue) and comfort (noise and
vibration). Traditionally, vortex induced vibrations meant
time-consuming trial-and-error searches for the exact source
of the vibration excitation, starting with “blind” modifications
of the most likely appendages as V-brackets, fins, sea chests,
etc. Menzel et al. (2008) show how modern simulations allow
efficient trouble-shooting. Initial measurements during sea
trials determined unacceptable vibration levels in the owner’s
Seakeeping
Fig.6
Seakeeping simulations checked loads on large window front for
cruise vessel
63
2.3
Aerodynamics flows
Aerodynamic flows around ship superstructures can be computed by CFD, in fact by the same code as used for slamming
and sloshing. CFD offers several advantages over wind tunnel
tests, perhaps most importantly not suffering from scale effects
which can be significant if thermodynamic processes are involved, El Moctar and Bertram (2002). Aerodynamic simulations
can be motivated by safety (e.g. safe landing of helicopters
on a helideck, Fig.10), fuel efficiency of fast ships, Schmode
and Bertram (2002), or comfort of passengers on deck, where
smoke dispersion is a key issue. Most recently, such investigations have been coupled to parametric modelling and formal
optimization, Harries and Vesting (2010).
Fig. 7
Snapshot of solar powered catamaran in waves
Fig. 8
Slamming simulation guiding design for better comfort on
megayacht
Fig. 9
Sloshing simulation guiding design for LNG tank (lighter
structure for given safety level)
Impact loads due to ship motions are another field where
simulations are used to assess safety and comfort aspects in
design. Slamming denotes external impact loads on ship hulls,
Fig.8, Köhlmoos and Bertram (2009). Sloshing denotes violent
fluid motion inside tanks, which pose particularly for LNG
(liquefied natural gas) carriers a problem, Fig.9, Schellin et al.
(2007). From a simulation point of view, both are impact loads
due to highly nonlinear free-surface flows. The same code is
employed with the same basic parameter setting.
64
Fig 10
CFD aerodynamic simulation to investigate safe helicopter landing
Fig.11
Funnel optimization for minimum smoke dispersion
3
Structural analysis
Fuel can be saved not only by hydrodynamic simulations.
At first glance, it is trivial: Smaller ships consume less fuel.
But the potential for weight reduction is rarely used to its full
extent. Progress in software tools allows today cost effective
weight savings through rapid modeling. Germanischer Lloyd’s
POSEIDON software is a structural product data modeler with
integrated finite element analysis (FEA) code, Fig.12. POSEIDON allows relatively fast modeling of the main ship structures.
Once the product data model is established, one can compare
scantlings automatically against Class Rules and identify where
ship structures are overdimensioned. But then more sophisticated FEA simulations should verify that no problems in terms of
vibration and fatigue are to be expected, i.e. that the increase
in inefficiency did not compromise accepted safety levels. In
a recent application, the initial steel weight of a multi-purpose
carrier was reduced by more than 2% checking first scantlings
with POSEIDON and then verifying critical spots by FEA. The
savings in steel cost alone would have justified the analysis for
a single ship. The savings are much larger when considering
series of ships and the accumulated fuel savings (which are
approximately proportional to the total weight of the ship).
a structural design with increased collision resistance could
eliminate the need for additional bulkheads, thus making the
ship cheaper to build and lighter (hence more fuel efficient).
Based on extensive FEA simulations for ship collisions,
Germanischer Lloyd developed an approval procedure which
provided the first such standard for evaluation and approval
of alternative solutions for design and construction of these
ships, Zhang et al. (2004), Fig.13.
4
Conclusions
The technological progress is rapid, both for hardware and
software. Simulations for numerous applications now often aid
decisions, sometimes ‘just’ for qualitative ranking of solutions,
sometimes for quantitative assessment and sometimes for
formal optimization. Frequently, several dedicated simulation
codes are used to improve efficiency or comfort aspects of a
design, while checking that sufficient safety is maintained.
Advanced designs are obtained through advanced software.
However, advanced software alone is not enough. Engineering is more than ever the art of modelling, finding the
right balance between level of detail and resources (time,
man-power). This modelling often requires intelligence and
considerable (collective) experience. The true value offered
by advanced service providers lies thus not in software licenses or hardware, but in the symbiosis of highly skilled staff
and these resources.
Fig. 12
Expert assessment of ship structure based on automatic rules checks
and selected simulations saved 200 t steel weight for this case
Acknowledgements
Many colleagues from the GL Group have supported this paper
with their special expertise, supplying text and/or figures,
namely (in alphabetical order) Christian Cabos, Bettar El
Moctar, Stefan Haries, Karsten Hochkirch, Jürgen Jokat, Axel
Köhlmoos, Holger Mumm, Tobias Zorn.
References
Fig.13
Finite-element simulation for collision analysis to prove equivalent
safety of lighter tanker design
Alternative design options in SOLAS allow some flexibility
of structural designs supported by advanced simulations. E.g.
ABT, C.; Harries, S. (2007), “A new approach to integration of
CAD and CFD for naval architects”, 6th Conf. Computer
and IT App lications in the Maritime Industries
(COMPIT), Cortona, pp.467-479.
http://www. ssi.tu-harburg.de/doc/compit/compit2007_
cortona.pdf
65
B ERTRAM , V. (2000), “Practical Ship Hydrodynamics”,
Butterworth & Heinemann, Oxford.
B ERTRAM , V.; Fach, K.; Sames, P.; Höppner, V. (2009),
“Engineering options to reduce emissions”, Int. Marine
Design Conf., Trondheim
ZIEGLER, W.; Fach, K.; Hoffmeister, H.; El Moctar, O.; Bethane,
P. (2006), “Advanced analyses for the EARTHRACE
project”, 5th Conf. High-Performance Marine Vehicles
(HIPER), Launceston, pp.101-108
EL MOCTAR, O.M.; Bertram, V. (2002), “Computation of viscous
flow around fast ship superstructures”, 24th Symp. Naval
Hydrodyn., Fukuoka, pp.68-77.
HANSEN, H.; Freund, M. (2010), “Assistance tools for operational
fuel efficiency”, 9th Conf. Computer and IT Applications in
the Maritime Industries (COMPIT), Gubbio, pp.356-366
http://www.ssi.tu-harburg.de/doc/webseiten_dokumente/
compit/dokumente/compit2010_gubbio.pdf
HARRIES, S.; Vesting, F. (2010), “Aerodynamic optimization of
superstructures and components”, 9th Int. Conf. Computer
and IT Applications in the Maritime Industries, Gubbio,
pp.335-347http://www.ssi.tu-harburg.de/doc/webseiten_
dokumente/compit/dokumente/compit2010_gubbio.pdf
KÖHLMOOS, A.; Bertram, V. (2009), “Simulation-based design of
super and mega yachts”, RINA Conf. Design, Construction
& Operation of Super & Mega Yachts, Genoa
MENZEL, W.; El Moctar, O.M.; Mumm, H. (2008), “Advanced
thinking on tricky excitations”, The Naval Architect, March,
pp.64-69.
OOSSANEN, P. Van; Heimann, J.; Henrichs, J.; Hochkirch, K.
(2009), “Motor yacht hull form design for the displacement
to semi-displacement speed range”, 10th Int. Conf. Fast Sea
Transportation (FAST), Athens
http://www.futureship.net/download/vanOossanen_etal
FAST2009_Paper143_090630.pdf
RATHJE, H.; Beiersdorf, C. (2005), “Decision support for
container ship operation in heavy seas - Shipboard Routing
Assistance”, 4th Int. Conf. Computer and IT Applications in
the Maritime Industries, Hamburg, pp.455-467
http://www.ssi.tu-harburg.de/doc/webseiten_dokumente/
compit/dokumente/compit2005_hamburg.pdf
SCHELLIN, T.; Peric, M.; El Moctar, O.; Kim, Y.S.; Zorn, T.
(2007), “Simulation of sloshing in LNG-tanks”, 26th Conf.
Offshore Mechanics and Arctic Engineering (OMAE), San
Diego
S CHNEEKLUTH , H.; Bertram, V. (1998), “Ship design for
efficiency and economy”, Butterworth & Heinemann,
Oxford.
S CHMODE , D.; Bertram, V. (2002), “Aerodynamic flow
computations for a Superfast ferry”, 3rd Int. Conf. HighPerformance Marine Vehicles (HIPER), Bergen, pp.345-354
66
USP wave basin: active wave absorption and generation algorithms
Mario L. Carneiro1, Pedro C. de Mello1, Felipe Labate1, André A. M. Araujo1, Alexandre N.
Simos2 and Eduardo A. Tannuri1
1 Dept. Mechatronics Engineering, USP
[email protected]
2 Dept. Naval Architecture and Ocean Engineering, USP
Numerical Offshore Tank - TPN-USP
Abstract
This paper presents the analysis of wave absorption and generation algorithms that were latter applied in the new wave basin constructed at the University of São Paulo (USP), as part of the Numerical Offshore Tank (TPN) Laboratory. These algorithms calculate
the motions of the wave makers both to generate and absorb the required wave field by taking into account the layout of the flaps
and the limits of wave generation. In order to study different aspects of the implementation, the performance of a prototype device
composed of 4 flaps was evaluated in a 2D wave flume, prior to the assembly of the complete system at the TPN wave basin. The
generation algorithms are based on the summation of wave components (frequencies and directions) obtained from the required
directional wave spectrum. The transfer function that relates the flap motion to the generated wave is considered. Absorption tests
were conducted using two different algorithms: a frequency domain method based on Maeda et al. (2004), in which the controlled
variable is the motor velocity, and the time domain algorithm proposed by Schäffer (2001). The latter is based on a digital filter and
the position of the flap is the variable to be controlled. Both algorithms require hydrodynamic feedback based on the measurement
of the surface elevation at each flap. The first algorithm needs an extensive test procedure to calibrate its control parameters, while
the second one, after optimizing the digital filter, should be ready to use. Both algorithms presented similar results with reflection
coefficients smaller than 10.7% for regular waves with frequencies ranging from 0.5 to 1.5 Hz.
Keywords
Wave Basin, Wave Absorption, Segmented Wave Maker, Reflection Coefficient, Digital Filter.
1
Introduction
This paper presents a detailed description of the algorithms development for the new active absorption wave basin built at the University of São Paulo (USP), a part of the Numerical Offshore Tank (TPN) Laboratory. The tank is intended to be used as a calibration
tool for the numerical models employed for the dynamic simulation of offshore structures and vessels. Furthermore, it may be used
as an efficient basin for testing the station keeping performance and motions of floating units in several kinds of wave fields. One
of the main goals pursued during its design was that the facility should have a simple and flexible operation and easy maintenance.
In order to achieve this feature, the wave basin was conceived to be small. Active wave absorption was then considered as a means
for reducing wave reflections.
The new facility is under development at the University of São Paulo since 2006. It consists of a 14m x 14m rectangular wave tank
with a depth of 4m and a wave generation-absorption system based on 148 flap-type wave makers. To develop the active absorption
system, each wave maker has an ultrasonic sensor that measures the instantaneous water level in its face. A high level control system
interface is present to allow the development of the control algorithms.
The system described by Salter (1981) was the first operational active absorption system in a multidirectional wave basin. Active
wave absorption has also been used in two newly opened facilities: The Amoeba (Advanced Multiple Organized Experimental
Basin) wave tank in Osaka (Naito et al. (1996) and Naito (2006)) is a prototype tank of variable geometry. Wave making is based
on a system of plungers and the wave absorption control is performed by monitoring the vertical velocity and the force on each
wave maker. A much larger facility was opened in 2002 at the National Maritime Research Institute (NMRI) in Tokyo. The Deep-Sea basin (Maeda et al (2004)) consists of a circular wave tank with a diameter of 16m, equipped with a set of 128 flap-type wave
makers along its circumference. The methodology for absorption control is different from the one employed at the Amoeba basin:
wave-probes mounted on each flap measure the wave elevation continuously and, by comparing it with predicted values, provide
the data necessary to correct the input signal for flap motion in order to absorb the reflected waves.
Submitted to MS&OT on Sep 09 2010. Revised version submitted on Nov 16 2010. Accepted on Dec 10 2010. Editor: Marcelo A. S. Neves.
67
Mario L. Carneiro, Pedro C. de Mello, Felipe Labate, André A. M. Araujo, Alexandre N. Simos and Eduardo A. Tannuri
This paper intends to describe the analysis, validation and
preliminary calibration of the generation and active absorption
algorithms. This work was executed in a 2D wave flume, that
was the workbench used for developing all the algorithms
later implemented in the TPN wave basin. Wave generation
is based on the time realization of frequency domain transfer
functions for each flap, considering linear superposition of
the individual responses. Preliminary tests conducted with a
small scale device were used to evaluate the performance of
the absorption algorithms, one of the main issues of the wave
tank. Two different algorithms were tested. The first one is the
frequency domain method based on Maeda et al. (2004) and
the second is the time domain algorithm proposed by Schäffer
(2001). Tests indicated the advantages of the second method
that was then in fact implemented in the wave tank. The commissioning and preliminary operation of the TPN wave basin
is described by de Mello et al. (2009).
2
Wave generation algorithms
The wave generation theory using flap-type wave makers is
well addressed in the specialized literature. In the linear theory
context, multi-directional waves can be generated by the summation of many wave components with different frequencies
and directions, as described by Nohara et al. (1996). The
summation in frequency can be made according to a prescribed power spectrum (S(w)), while the summation in direction
follows a energy spreading function (D(θ)). The final result
is a short-crested wave field, as indicated in eqs. 2.1 and 2.2.
When the summation is made only in frequency, the result is
a long-crested sea (D(θ) is zero for every direction other than
the prescribed wave direction). Regular waves are obtained by
considering only one amplitude and one frequency.
number of active flaps depends on the desired ranges of wave
frequencies and directions to be generated (de Mello et al,
2009). The TPN wave basin is composed of 148 active flaps
(two of its sides are equipped with 39 flaps and the other
two have 35 active flaps each). Therefore, the generation of
oblique waves by two adjacent sides requires the activation
of 74 flaps. In the same way, when directional spreading is
considered, the number of directions increases and more flaps
must be activated to generate the desired wave. For example,
for delivering a short-crested wave with 0º mean direction and
energy spreading from -180º to 180º, three sides of the wave
basin must be used (109 flaps).
The wave generation algorithm was implemented using
MATLAB®. Concerning the computation effort required for
processing the generation data, the regular wave takes short
time to be executed, while the long-crested wave has an intermediate computational cost and the short-crested wave is quite
demanding. The calculation of each flap motion is done by a
pre-processor and the time series are generated and loaded by
the wave generator real-time control software. Table 1 provides
a comparative example of the computational cost required to
compute the time series for different wave fields. All cases
consider wave generation for a period of 120 seconds, sample
rate of 83Hz, two mean directions, 540 frequency components
and 36 direction components when applicable.
Table 1 Example of the computational cost required to
calculate the time series
(2.1)
where anm is the wave amplitude of the component of frequency
n, and direction m, given by:
(2.2)
and:
Xi : i-th wave maker stroke (position)
N : Number of frequency components
M : Number of directional components
I : Number of wave makers
Fn : n-th frequency 2D flap transfer function for the flap
at driven height
fn : n-th wave frequency
: Wave maker width
θm : m-th wave direction
kn : n-th wave number
εnm : Random phase in the interval [0...2π]
S(wn ) : Power spectrum (m²s)
D(θm ) : Energy spread function with
3
Wave generation transfer
function
The wave generation transfer function relates the progressive
wave amplitude (AI) generated in the tank with the amplitude
of the sinusoidal motion (X0) of each flap, for different wave
frequencies. It is discussed, for example, in Dean and Dalrymple
(1984) or Schäffer (1996). At a distance far enough from the
wave generator, the transfer function can be written as:
(3.1)
where i is the imaginary number and represents the 90o phase
shift between the wave and the flap movement, θ is the wave
propagation direction. For a flap type wave generator, c0 can
be written as:
The transfer function Fn that relates the flap stroke to the wave
amplitude is discussed in the next section.
Considering the geometry of the wave basin, the required
68
(3.2)
where kp is the wave number of the progressive wave, h is the
depth of the tank, h1 is the vertical distance from the tank bottom
to the flap pivot point and h2 is the flap total height.
Closer to the wave maker, localized effects (also known as
evanescent wave modes) are observed:
(3.3)
where A0 is the complex amplitude of the wave level measured
in front of the flap and ej is the transfer function of the j-th
evanescent wave mode, which, for a flap type wave generator,
is given by:
also included, in order to avoid drifting of the flap, keeping its
motion around the neutral position:
(4.1)
In the relation above, vk is the imposed velocity of the flap, k
is the number of the flap, xk is the flap position with respect to
its neutral position and K is an adjustable control gain (used
to avoid the drift). The component responsible for the wave
absorption was simplified, considering a perpendicular incidence angle:
(4.2)
where fp is the instantaneous estimated wave frequency, Fp is
the progressive transfer function (equal to c0) and ηrk is the
reflected wave elevation.
(3.4)
in which kj is the wave number of the evanescent mode j and
kxj is its x-component.
An example comparing the progressive wave transfer function
AI / X0 (equation 3.1) and the full wave generation transfer
function A0 / X0 that includes progressive and evanescent modes
(equation 3.3) is shown in Figure 1:
During the preliminary tests in a 2D wave flume (described
in the next section), high-frequency oscillations of the flaps
were observed. A possible explanation for such fact is the
influence of evanescent modes in the measured wave height,
which were neglected during control derivation. A correction
was then introduced in the reflected wave elevation measurement (Kawaguchi, 1986), considering that evanescent
wave modes and flap acceleration present approximately no
). Furthermore, a reduction in the
phase shift (
overall gain of the controller was also included (Kp). The final
formulation for the reflected wave elevation is then given by:
(4.3)
where ηk is the wave elevation measured in the flap k and
ηdk is the reference wave elevation, previously calculated by
equation 3.1.
The instantaneous frequency of reflected wave is obtained from
standard frequency estimation algorithms. In the present paper,
the following relation was used:
Fig. 1
4
(4.4)
Wave generation transfer functions
5
Absorption algorithm A
The first absorption algorithm tested is based on the one proposed by Maeda (2004). This algorithm uses hydrodynamic
feedback based on the measurement of wave elevation at each
flap and flap velocity as a reference for the control.
Neglecting the evanescent wave modes, it can be shown that
there is no phase shift between wave height and flap velocity
and this justifies the use of a reference velocity signal, since a
non-complex (pure-real) control gain can be used.
The actuation signal is composed by two terms. The first one
is the flap velocity required to generate the desired wave vdk,
previously calculated by the generation transfer function (equation 3.1). The second term is responsible for the absorption
of reflected waves vak . In the present work, a third term was
Absorption algorithm B
The second algorithm is based on Schäffer (2001). Similar to
the first algorithm, it also uses the free surface elevation in front
of each wave maker as hydrodynamic feedback, but here the
position reference is used as control signal.
This algorithm is developed in time domain and includes the
effects of the evanescent modes. Considering a perpendicular incidence angle, the wavemaker position X0 can be formulated as:
(5.1)
(5.2)
where ‘*’ denotes complex conjugate. AI is the complex amplitude of incident wave ηdk , F0 is a complex transfer function
69
related to the inverse of the flap generation transfer function
(A0 /X0) and its implementation in time domain is made by a
recursive digital filter:
The delay originated from the position loop motor control can
be compensated by optimizing F0M = F0/M instead of the original absorption transfer function F0. The final block diagram of
the absorption control system is showed in Figure 3.
(5.3)
The coefficients ak and bk can be obtained by optimization
to match F̃ and F0, remembering that the poles of the digital
filter must be within the unit circle in the z-plane to guarantee
stability. During the optimization, high frequency responses
should also be forced down to avoid instability, as shown, for
example, in Figure 2.
Fig. 3
6
Diagram of the control absorption algorithm B.
Experimental setup
Both absorption algorithms were tested in a wave flume at the
Naval Arch. & Ocean Eng. Dept. at Escola Politécnica. The
wave flume is 25m long, 1.0m wide, the still water level is 0.8m
and it is equipped with an edge type wave maker.
To absorb the waves, a prototype wave generator composed
of four independent flaps (Figure 4) was installed on the opposite side of the original wave maker. Each flap is equipped
with an ultrasonic wave sensor mounted on its face, based on
the propagation time of pulse-echo, as described by Martins
et al. (2007).
(a) Full frequency range
Fig. 4
Prototype absorbing wavemaker.
The experiments consisted of testing the absorption of regular
waves with different amplitudes and frequencies generated by
the original wave making system. The reflection coefficient (Cr
= Ar/AI) was estimated by the method proposed by Mansard &
Funke (Isaacson, 1991) and implemented by de Mello (2006),
using the signals gathered by an array of wave probes installed
in the center of the wave flume, as shown in Figure 5.
(b) Desired frequency range
Fig. 2
70
Example of filter optimization , where F0 is the target filter,
F0 A was optimized considering high frequency response and
F0 B was optimized using only the desired frequency range.
Fig. 5
Experimental setup
7
Results
In the present section, the results obtained with the absorption
algorithm in the wave flume are presented and discussed.
Figure 6A shows an example of wave elevation measured by
the wave probe installed on the absorbing flaps when they are
inactive. The progressive wave height generated by the original
wave-maker system is approximately 15 mm. It can be seen
that, due to the reflection in the inactive flaps, the elevation is
almost doubled as soon as it reaches the flap (during the first
50s), since the sensor measures the sum of incident and the
reflected waves. After that interval, the reflected waves reached
the wave-making system, reflected on it, and returned to the
flaps. This second reflection occurred in approximately 50~60s,
and can be clearly noticed in the figure.
Figure 7 presents an example of the time series of the reflection
coefficient Cr, incident and reflected wave heights estimated for
the two different absorption algorithms. This test corresponds to
a wave frequency of 1.0 Hz. The average reflection coefficient
during the interval between 80s and 140s is 6.7% for algorithm
A and 7.9% for algorithm B.
When the flaps are active, on the other hand, the absorption
algorithm avoids the reflection of a major part of the incident
wave. Figure 6B shows the replication of the previous experiment, now with the flaps operating. The efficiency of the method
can be readily noticed. In this case, the wave amplitude remains
approximately constant around 15~17mm.
In order to optimize the control parameters of the absorption
algorithm A, a trial and error fine-tuning procedure was used
for each case to obtain the smallest reflection coefficient. Only
the best result is presented for each generated wave.
(a) Algorithm A
(a) Absorption inactive
(b) Algorithm B
(b) Absorption active
Fig. 6
Example of generated waves.
Fig. 7
Example of experimental reflection coefficient (1.0 Hz wave).
71
In order to absorb irregular waves, estimation of the instantaneous wave frequency fp is necessary at each time step. This
estimation is applied to correct the control parameters K, Kp
and Kacc. During the tests with regular waves, however, it was
observed that the frequency estimation by means of equation
4.4 is very sensible to level sensor noise. A long averaging mean
filter with period of 1.0s was required to obtain a reasonable
result, consequently delaying the frequency estimation response. Figure 8 shows an example of the estimated frequency
before and after the filtering and the correspondent elevation
time series for a regular wave of frequency 1.0 Hz.
During the tests with the absorption algorithm B, high frequency
oscillations were also observed, than a reduction in the overall
gain of the controller was also included (Kp). This is the only
parameter to be adjusted. The summary results can be seen in
Table 3:
Table 3 Summary of experimental results – Absorption
algorithm B
8
Conclusions
In the present paper, a discussion about the wave generation
and absorption algorithms to be implemented in the TPN wave
basin was presented.
(a) Frequency estimation
Generation is based on the time realization of the frequency
domain transfer functions of each flap, considering their linear
superposition. Generation algorithms for regular and irregular
waves, with or without directional spreading, were studied, but
experimental validation still needs to be conducted.
Two different absorption algorithms were tested using a small
scale prototype in a 2D wave flume. Results indicated that:
l
l
Both algorithms presented acceptable performance for
regular waves, with reflection coefficients smaller than
11% for wave frequencies between 0.5Hz and 1.5Hz.
Algorithm A requires a time-consuming tuning process,
in order to adjust at least two control parameters. For each
wave frequency, a new tuning procedure is required.
(b) Elevation time series
l
Fig. 8
Example of frequency estimation.
Table 2 shows a summary of the results for the absorption algorithm A. It can be seen that the reflection coefficient was smaller
than 11% for all cases tested, what is considered quite acceptable.
Control parameters adjusted for each case are also presented.
l
l
Table 2 Summary of experimental results – Absorption
algorithm A
l
Algorithm A is executed in frequency domain and
an online algorithm for wave frequency estimation
is also required when considering irregular waves.
Algorithm B has only one tuning parameter (control gain Kp),
that is theoretically equal to 1. A reduction in this control gain
was introduced in order to avoid high frequency oscillations.
Algorithm B is executed in time domain, and no frequency
estimation is required.
Convergence, stability and tuning of the frequency estimation algorithm are yet to be evaluated, and may pose another
drawback concerning the application of Algorithm A.
Based on the considerations above, algorithm B, which was
based on the original method proposed by Schäffer (2001),
was chosen for the control system to be implemented on the
TPN wave basin.
72
Acknowledgments
The authors acknowledge Petrobras for the financial support
and for the motivation of this work. The first author acknowledges the São Paulo State Research Foundation (FAPESP
Proc. No. 2008/06428-4). Fifth and sixth authors also acknowledge the Brazilian National Research Council (CNPq) for
the research grants.
NOHARA, B. T.; Yamamoto, I.; Matsuura, M. (1996), “The organized motion control of multi-directional wave maker”,
Proceedings of the 4th International Workshop on Advanced
Motion Control, v. 2, p. 470-475.
OCHI, M. (1998), “Ocean waves. The stochastic approach”,
Cambridge Ocean Tech. Series 6, Cambridge Univ. Press.
SALTER, S. H.(1981), “Absorbing wave-makers and wide
tanks”, Proceedings Directional Wave Spectra Applications,
Berkeley.
SCHÄFFER, H. A. (1996), “Second-order wavemaker theory
for irregular waves”, Ocean Engineering, Vol. 23, No. 1,
pp. 47-88.
References
DEAN, R. G., and Dalrymple, R. A. (1984), “Water wave mechanics for engineers and scientists”, Prentice-Hall, Inc.,
Englewood Cliffs, New Jersey.
SCHÄFFER, H. A., 2001, “Active wave absorption in flumes and
3D basins”, Waves’01: Proc 4th Int. Symp. on Ocean Wave
Measurement and Analysis, San Francisco, USA, ASCE,
pp. 1200-1208.
ISAACSON, M. (1991), “Measurement of regular wave reflection,” Journal of Waterway, Port, Coastal, and Ocean
Engineering, Vol. 117, No. 6, pp. 553-569.
KAWAGUCHI, T. (1986), “Absorbing wave making system with
wave sensor and velocity control,” Mitsui Zosen Technical
review, No. 128, pp. 20-24, (in Japanese).
MAEDA, K., Hosotani, N., Tamura, K., and Ando, H. (2004),
“Wave making properties of circular basin”, International
Symposium on Underwater Technology, pp. 349-354.
MARTINS, J. A. de A., de Mello, P. C., Carneiro, M. L., Souza,
C. A. G. F., and Adamowski, J. C.(2007), “Laboratory wave
probes dynamic performance evaluation”, Proceedings of
XX COPINAVAL - Congresso PanAmericano de Engenharia Naval e Transportes Marítimos, São Paulo, Brazil.
DE MELLO, P. C. (2006), “Reduction of reflected waves in wave
tank with parabolic beach”, M.Sc. Dissertation, COPPE,
Federal University of Rio de Janeiro, Rio de Janeiro.
Available at: http://www.oceanica.ufrj.br/intranet/modules/
rmdp/down. php?id=35 (in Portuguese).
DE
MELLO, P. C., Carneiro, M. L., Tannuri E. A., Nishimoto
K. (2009), “USP active absorption wave basin: from
the conception to the commissioning”, 4rd International
Workshop on Applied Offshore Hydrodynamics, Rio de
Janeiro, Brazil.
NAITO, S., Nakamura, T., Sakashita, H., and Tomita, K. (1996),
“A new configuration of wave basin and a control of wave
generation and absorption-the case when an advancing
ship comes across the given waves”, Proceedings of the
4th Pacific/Asia Offshore Mechanics Symposium, Vol. 226,
pp. 207-212.
NAITO, S. (2006), “Wave generation and absorption in wave
basins: theory and application”, Proceedings 16th International Offshore and Polar Engineering Conference, San
Francisco, California, USA.
73
A
Numerical and experimental procedure for designing sub-sea installation operations
André L. C. Fujarra1, Eduardo A. Tannuri2, Felipe R. Pereira3, Rafael M. L. Madureira4,
Isaias Q. Masetti5 and Haroldo Igreja6
1 Department of Naval Arch. & Ocean Eng., Numerical Offshore Tank (TPN-USP), University of São Paulo, Brazil,
[email protected]
2 Department of Mechatronics Eng., Numerical Offshore Tank (TPN-USP), University of São Paulo, Brazil, [email protected]
3 Department of Naval Arch. & Ocean Eng., Numerical Offshore Tank (TPN-USP), University of São Paulo, Brazil,
[email protected]
4 Petrobras Transportes SA, Petrobras, Brazil, [email protected]
5 Petrobras Transportes SA, Petrobras, Brazil, [email protected]
6 E&P - Serv, Petrobras, Brazil, [email protected]
Abstract
Sub-sea equipment installations are very complex operations, requiring pre-installation analysis to define the correct procedure and
the weather “window” for a safe operation. This paper addresses the installation of a Mid Water Arch (MWA) intended to provide
support to the riser. Connecting the riser to the MWA largely eliminates the dynamic forces that would otherwise cause friction
and fatigue. The MWA is composed of riser guides and several buoyancy tanks and is kept in the water with tethers connected to
an anchor. The installation procedure involves launching each component of the MWA (anchor, main structure and tethers), during
which a tug boat with an A-frame and an assistance vessel are used to keep the buoy away from the tether and the launch cable. The
waves induce oscillatory motions throughout the system and may cause large dynamic forces in the cables and tethers. Due to the
complexity of the multi-body system, a comprehensive numerical and small-scale experimental analysis is conducted to calculate
the proper dimensions for the launch cables and to define the limits of the environmental conditions. Numerical analysis was carried
out in the Numerical Offshore Tank – TPN, a multi-processor offshore system simulator that considers the 6 degrees of freedom
for each body and all environmental forces acting upon them. The lines are modeled by finite-element analysis. Furthermore, a
full set of small-scale experiments were carried out at a towing tank that considered the response of the system when excited by
sinusoidal motion at the top and emulated the wave excitation. Comparisons between numerical and experimental results showed
good adherence between the calculated values. The validated numerical simulator was then used to analyze the complete complex
installation procedure by considering an extensive set of environmental conditions.
Keywords
Subsea installation; Simulation; Offshore operation
Nomenclature
BTA
Buoyancy Tank Assembly
CG
Center of Gravity
DOF
Degree of Freedom
IPT
State of São Paulo Technological Research Institute
LVDT
Linear variable differential transformer
MWA
Mid Water Arch
TPN
Numerical Offshore Tank
RAO
Response Amplitude Operator
ROV
Remotely Operated Vehicle
Submitted to MS&OT on May 24 2010. Revised version submitted on Nov 18 2010. Accepted on Dec 10 2010. Editor: Celso Pesce.
75
André L. C. Fujarra, Eduardo A. Tannuri, Felipe R. Pereira, Rafael M. L. Madureira, Isaias Q. Masetti and Haroldo Igreja
1
Introduction
Subsea equipments such as manifolds and riser supporting
systems require complex offshore operations to be launched
and positioned in the correct location in the seabed. Rowe et al.
(2001) presented several problems associated with subsea launching. Although the focus of that work is deepwater operations,
the main concerns may be also extrapolated for all launching operations. The most relevant problems pointed by the authors are:
• Lifting equipments: problems associated with the loads to
be lowered, dynamic amplification during the launch and
capability of the equipments;
• Load control and positioning: problems associated with the
correct final laying positioning of the subsea equipment;
• Weather conditions: problems associated with launching vessel induced motions and weather window to a safe operation.
The present paper presents a methodology to analyze complex
offshore operations involving sub-sea installations and several
support vessels. The execution of full-scale experiments involving all vessels and components of the actual operation may be
extremely complex and expensive. Furthermore, depending on the
needs of the offshore industry, the time required to prepare and
execute such experiments may not make the experiments feasible.
search Institute – IPT towing tank. These results were then
used to validate a numerical model developed at TPN from the
University of São Paulo. Comparisons between the numerical
and small-scale experimental results indicated that the numerical
model was a reliable tool to predict the system’s behavior during
installation. Complementary numerical simulations were completed to consider extreme wave conditions and an irregular sea
spectrum. The simulations indicated some operational problems
that may occur during the installation, and the results were used
to re-design specific steps of the procedure.
The main contribution of the paper is to present a successful
case study of the hybrid methodology (simplified experiments
and full time domain numerical analysis) applied to a complex
subsea launching operation, under environmental conditions.
Particularly, two critical steps of the launching procedure were
addressed. Case 1 was defined as the step during which the
anchor is lowered through the water while the tethers remain
tension free since the BTA sits on the water’s surface. It is important to clarify that the anchor is supported by a lifting wire
that is connected to the A-frame of the main installation vessel
and the BTA is kept away from the lifting wire by the assistance
vessel (see Fig. 1). Case 2 differs in that tension is added to the
tethers, which pull the BTA into the water. Here again, the BTA
is continuously kept away from the lifting wire by the assistance
vessel, as seen in Fig. 2.
Numerical simulation is a tool that engineers use for performing
analysis prior to actual installation. Ferreira (2002) presented
an extensive numerical analysis of a conventional manifold
installation procedure using a linear frequency domain analysis.
In that work, the importance of the coupled dynamic analysis
was stressed. An alternative launching method using two vessels
was presented by Santos et al, (2009). Nonlinear time domain
simulation were used for predicting the loads in the cables, but no
dynamic coupling between vessels and the load has been considered. In those works, no experimental validations were presented.
However, due to the enormous complexity of some operations,
the engineers cannot rely only on the numerical simulation results
to make important decisions. A combination of numerical and
experimental analysis was presented by Fernandes et al. (2006)
for the evaluation of the pendulum method for subsea launching.
This launching procedure requires one vessel and no environmental condition is considered. In that case, fundamental aspects
of the experimental results were recovered by simulations, but
rotational motions of the manifold could not be predicted. The
results obtained in the analysis were important for the definition
of the real operational installation procedure (Lima et al. 2008).
Fig. 1
The launch of the anchor - Case 1
Fig. 2
The launch of the anchor connected to the BTA - Case 2
This paper presents an example of a hybrid methodology to
analyze a complex sub-sea equipment installation. Simplified
experiments were used to validate the numerical simulations,
which were then used for further complex simulations for the
full-scale operation under real environmental conditions. The
installation of a Mid Water Arch (MWA) is considered. The
MWA is a structure that provides riser support and consists of
a Buoyancy Tank Assembly (BTA), an anchor and two tethers
connecting the BTA to the anchor. During the installation, supplementary cables and two tug boats are employed.
A full set of simplified and low-cost, small-scale experiments
were carried out in the State of São Paulo Technological Re-
76
2
Table 2 BTA properties
Modeling data
As previously mentioned, each MWA consists of a BTA, an
anchor and two tethers connecting the BTA and the anchor.
During the installation, supplementary cables and two tug
boats are used. The technical descriptions of the equipment
components and vessels involved in the operation are presented
in this section. These data were used in the implementation of
the numerical modeling of the operations and the experimental
set-up.
The anchor is assumed to be of gravity-type and consist of a
steel-reinforced concrete slab. It is equipped with two tether
connection lugs and four lifting lugs. The main weight properties of the anchor are summarized in Table 1, and a principle
sketch with the main dimensions is shown in Fig. 3.
Table 1 Anchor properties
Fig. 3
Fig. 4
Pictures of the BTA
Fig. 5
MWA model and cable system
Anchor and cable assembly
The BTA consists of two buoyancy tanks, a riser installation
guide, a main frame steel structure, two hinged frames for tether
connection and four lifting lugs for tug connection. Figure 4
presents images of the BTA during the construction, and Table
2 presents the main dimensions and weight properties of the
BTA. The cables that connect the BTA to the assistance vessel
during the installation of the MWA are shown in Fig. 5. Tethers
of studless chain, 42m in length and 76mm in diameter, connect
the anchor to the BTA. Each one weighs 48kN and presents a
breaking load of 5,448kN approximately.
A lifting wire (launch cable) connects the anchor to the main
vessel during the installation of the system. It is an 84-mm diameter steel cable with a breaking load of 4,312kN and stiffness
equal to 216,300 kN/m². Two tethers connect the anchor to
the BTA. They are 76-mm diameter studless chain, with 42m
length and submerged weight of 44kN each.
The main installation vessel is considered to be similar to the
Normand Neptun tug boat, shown in Fig. 6, and is equipped
77
with an A-frame for launching the anchor. According to Fig.
7, by assuming that the A-frame connection point is located
12.2m above and 47.2m from the CG, the Response Amplitude Operators (RAO’s) of the vessel were obtained using
the software WAMIT (WAMIT, 2000). Figure 8 presents the
A-frame point heave and surge RAO for a head-sea incident
wave. Large amplification of heave motion (factor 2.1) for
wave periods close to 8s was verified.
Fig. 6
Picture and main characteristics of the Normand Neptun
Fig. 7
Position of the A-frame in the Normand Neptun
Fig. 8
The RAO evaluated at the A-frame point (in head-sea waves)
Head-sea waves, with 2.0m height and 9.0s peak period, were
considered in the numerical and experimental analysis, for the
installation location of 110m in depth.
Case 1 comprises when the anchor is being lowered through the
water, and the tethers are tension free because the BTA is still
on water surface. A simple static analysis revealed that the mean
traction in the lifting wire would be approximately 774kN, as
presented in the Fig. 10. The transition to Case 2 occurs when
the tethers are under tension, while the anchor is being lowered,
which pulls the BTA into the water. In both cases, the BTA is
distanced from the lifting wire by the assistance vessel.
Fig. 10
Static analysis of Case 1
The static analysis was performed considering two different
values for the horizontal tug force of the assistance vessel:
98kN and 196.2kN (10 and 20 ton, approximately). For the
98kN case, Fig. 11 shows the mean forces on each cable, and
Fig. 12 shows the final static configuration. For the 196.2kN
case, the static forces are presented in Fig. 13, and the geometrical configuration is shown in Fig. 14.
A vessel similar to the Sidney Candies tug boat has also been
considered for use as the assistance installation vessel. The
main characteristics of this vessel are presented in Fig. 9.
Fig. 9
78
Picture and main characteristics of the Sidney Candies
Fig. 11
Static analysis of Case 2 (98kN auxiliary tug force)
The mean traction in the lifting wire increased from 338kN
to 386kN when the tug force was increased from 98kN to
196.2kN, which reduced the occurrence of cable slackening
during launching, as will be discussed later.
3
Experimental details
Tests were conducted at the State of São Paulo Technological
Research Institute –IPT towing tank. According to the Petrobras
requirements for design, four test conditions were applied:
• Case 1 - the launch of the anchor;
• Case 2 - the launch of the anchor connected to the BTA;
Fig. 12
Static analysis of Case 2 (98kN auxiliary tug force) – final configuration
Fig. 13
Static analysis of Case 2 (196.2kN auxiliary tug force)
Considering the towing tank dimensions (6m wide and 4m deep),
a scale of 1:50 was chosen for modeling the anchor, BTA and mooring system. Fig. 15 presents some photos of the anchor and BTA
models. This small scale may be used to extrapolate results to full
scale. The flow separation points around the anchor are well defined
(since the anchor is a sharp-edge box). Therefore, the anchor drag
force is weakly dependent on the Reynolds number, and is well
predicted in the experiments. Furthermore, the most important
dynamical effect that arises during the launching (cable slackening)
is directly affected by the anchor (and not by the BTA) drag.
Fig. 15
Fig. 14
Static analysis of Case 2 (196.2kN auxiliary tug force) – final configuration
Small-scale images of the anchor and BTA models (scale 1:50)
The dynamics of the tug boat during the launch were considered
as equivalent vertical movements imposed by means of a servo-controlled linear actuator (see Fig. 16). Sinusoidal movements
were selected and applied during the vertical movements, according
to the combinations of amplitude and periods shown in the Table 3.
79
mic amplification factors (defined as the relation between the
maximum force and the mean force).
Fig. 16
Images of the servo actuator and the table for anchor landing
Table 3
Amplitude and period (Ai;Ti) combinations adopted
for the sinusoidal movements imposed to the top
of the launch cable. Values presented in full scale.
Fig. 17
Monitored anchor during the launch test – Case 1
Table 4 Results of force on the anchor (Case 1 tests).
In Case 1, nine sinusoidal movements were imposed to two
different lengths of the lifting wire inside the water, L = 20m
and L = 60m (dimensions in full scale). In Case 2, the anchor
connected to the BTA was tested for L = 60m and L = 100m.
In order to emulate ground effects, a submerged table was
constructed for the last depth, also shown in Fig. 16. It is important to emphasize that a grid was applied to the submerged
table in order to evaluate the azimuth of the anchor landing.
Furthermore, for Case 2 conditions, two different angles between the launch line of the anchor and the tethers connecting
to the BTA were considered: approximately 28° and 54°.
A load cell with fine resolution mounted on the top of the launch
cable measured the time-varying traction. A LVDT was used
to measure the sinusoidal movement imposed to the top end
of the launch line, and two biaxial accelerometers were integrated into an internal compartment of the anchor. Tests were
filmed by a set of two cameras positioned above and lateral
to the experimental setup, respectively. The first camera was
installed at the carriage, and the second camera filmed through
the inspection window of the towing tank.
In Fig. 18, the dynamic amplification factors are presented as a
function of the imposed motion amplitude. As expected, when
the sinusoidal amplitude increases, or the period decreases, the
dynamic amplification factor increases.
4
Experimental results
4.1
Case 1 - The launch of the anchor
First, the launch of the anchor with an imposed oscillatory
movement to the top end of the launch line was considered.
A photo of the anchor during the launch is shown in Fig. 17.
Table 4 presents the mean, maximum and minimum forces
obtained from anchor launch, as well as the respective dyna-
80
According to the graphs in the Fig. 19, no combination exhibited a zero traction value in the analyzed amplitude and period
ranges. The maximum force was 1285kN for 3m amplitude and
5.89s period. As a general procedure, all the time histories of
force were pre-filtered in order to achieve the maximum, mean
and minimum values. The pre-filtering was necessary due to
high frequencies present in the signals. The signals were most
likely associated to oscillations of launch line, which was not
in the same scale as the anchor and BTA models. Figure 20
compares the original and filtered time histories of force for an
imposed sinusoidal movement with an amplitude and period
of 1.50m and 8.83s, respectively.
Fig. 20
Fig. 18
Dynamic amplification factor of the force measured from above during
the anchor launch
Example of the force measurement for the combination MAPH3008A:
the original time history and the filtered signal
Due to the heave amplification in the A-frame point (Fig. 8), a
2.00m incident wave with period 8.83s corresponds to 3.90m
motion amplitude in the lifting wire. Consequently, numerical
simulations (that will be detailed later) indicated a higher value
of force as compared to those shown in the combinations of
Table 4 combinations. In fact, the simulations of a 2.00m incident irregular wave with period 8.83s induced a maximum force
of 1100kN (see Table 6). It is interesting to note that a linear
extrapolation of the experimental results indicated a 1055kN of
maximum force. This is consistent with that obtained in TPN
simulations (see Fig. 21). Further details about the numerical
simulations are found in the next section.
Fig. 19
Forces at the launch line during anchor installation
Fig. 21
Maximum forces at the launch line as function of the amplitude for the
imposed sinusoidal movement. Linear extrapolation for A = 4.2m and
T = 8.83s.
4.2
connected to the BTA
Tests were completed to examine the effects of connecting the
BTA to the anchor for two different lengths of the lifting wire,
60m and 100m. The same sinusoidal movements as in Case 1
were applied to the launch line. At 60m, two static configurations were considered. The first one corresponded to an angle
of approximately 28° between the launch line and the tendons
connecting the BTA. These conditions were equivalent to
81
Faux=98kN on the line from the BTA to the assistance installation vessel. In the second configuration, Faux=196.2kN, was
applied to the line connecting the BTA and the tug boat. This
configuration corresponded to a 54° angle. Both geometries
are presented in Fig. 22. However, only the 28° configuration
was tested at 100m.
Fig. 22
Table 5 Results of force on the Case 2 tests.
Static configurations for Case 2 - the launch of the anchor connected to the BTA:
angles of approximately 28° and 54°between the tendons and the l launch line.
Table 5 presents the results of all experimental configurations
for Case 2. For each test, launch forces and dynamic amplification factors were obtained by the previous procedure
(maximum, mean and minimum values).
Figures 23, 24 and 25 depict the forces and dynamic amplification factors as a function of the sinusoidal amplitude and
the traction at the assistance installation vessel (Faux=98kN
and Faux=196.2kN, respectively). The graphs consider both
configurations at L=60m, as well as the three periods of movement. As expected, as the amplitude of movement increases,
the associated dynamic forces and amplification factor are
amplified. The higher forces were observed for 5.89s period.
82
Fig. 23 (a) Dynamic amplification factor and
(b) Values of the force at the anchor launch line with length L = 60m
for sinusoidal movements with period T = 8.83s.
Additionally, it was verified that by increasing the assistance
installation vessel force, a reduction of the dynamic amplification factor could be observed. Such a behavior can be better
distinguished by comparing the graphs in Fig. 26, where force
time histories at the launch line are presented for both cases of
auxiliary tug force (98kN and 196.2kN, respectively) at two
different periods of imposed movement.
Figures 27, 28 and 29 present the cases of an auxiliary tug
force of 98kN, at two values of lifting wire length (L), 60m
and 100m. As the depth increased there was a reasonable amplification of the traction at the launch line of the anchor. The
same conclusion could be obtained though observation of Fig.
30, which compares the tests from different periods.
Fig. 24
(a) Dynamic amplification factor and (b) Values of force at the anchor
launch line with length L = 60m for sinusoidal movements with a period
of T = 5.89s.
Fig. 25 - (a) Dynamic amplification factor and (b) Values of force at the launch line
of anchor with length L = 60m for the sinusoidal movements with a period
of T = 11.79s.
Fig. 26
Time histories of force at the launch line for a 2.25-m amplitude and lifting
line length of 60m.
83
Fig. 27
Fig. 29
launch line for period T = 11.79s, Faux = 98kN and two values of line
length (60m and 100m).
Fig. 30
Dynamic amplification factor as a function of the period, A = 2.25m and
Faux = 98kN.
launch line, for period T = 8.83s, Faux = 98kN and two values of line
5
Numerical analysis and validation
Numerical models of the MWA launching procedure were
programmed in the Numerical Offshore Tank – TPN. The TPN
is a multi-processor offshore system simulator that considers
the 6DOF for each body and all environmental forces acting in
them, as well as complex finite element models for the cables
and mooring lines (Nishimoto et al, 2003). A full description
of the models included in TPN is given in the Appendix.
Fig. 28
84
(a) Dynamic amplification factor and (b) Values of force at the launch
line of anchor for period T = 5.89s, Faux = 98kN and two values of line
Simulations with the same conditions as the experimental cases
were carried out, in order to validate the numerical models with
the pre-existing small-scale experiments. After comparable results were verified between both methods, simulations involving
irregular waves were performed. This allowed for evaluation
of the behavior of the actual system during offshore operation.
5.1
Irregular waves - real system behavior
The launch of the anchor was simulated with several differing
lengths of lifting wire. An example of the 3-D view of the
numerical model is show in Fig. 31.
The behavior of the system under a real sea state was predicted
by means of numerical simulations. The simulations considered several lifting wire lengths and wave significant height of
2.00m with a 9.00s peak period (Pierson-Moskowitz spectrum).
Table 6 and Fig. 33 show these results. The wire traction was
smaller than 1100kN for all cases without the occurrence of
cable slackening. Each simulation considers 500s of operation.
Table 6 Lifting wire traction, irregular waves
(Tp = 9.0s; Hs = 2.0m) – Case 1.
Fig. 31
TPN model for the launch of the anchor (10m lifting wire).
Regular waves - comparison with experiments
Numerical simulations with regular waves were carried out,
to allow for direct comparisons with the experimental results
presented in the previous section. Figure 32 shows the results
when the wave period was defined as 8.83s. A very good
adherence between numerical and experimental results was
verified for all amplitudes considered.
Fig. 33
5.2
Irregular waves (T p = 9.0s; Hs = 2.0m) - Case 1.
connected to the BTA
Several numerical simulations were carried out with the anchor directly connected to the BTA during launch in order to
make comparisons with experimental results and to predict
the behavior of the system under real sea conditions. The 3-D
view of the numerical model is shown in Fig. 34.
Fig. 32
Regular waves (T = 8.83s) - Case 1 - comparison between experimental
and numerical simulation results.
Fig. 34
TPN model for the launch of the anchor connected to the BTA - Case 2.
85
Regular waves - comparison with experiments
Table 7 shows the results of several numerical simulations
of Case 2, considering regular waves as compared to experimental results. Such a comparison is also presented in
Fig. 35. An acceptable adherence between the results was
seen, except when the wave period was 5.90s, in which
case the dynamic amplification factors obtained in the
experiments were higher than those obtained in the numerical simulation. Such difference may be explained by the
cable dynamics, since the stiffness used in the experiments
was not exactly the same used in the simulations, due to
practical limitations.
It is worth mentioning that for the situations with greater
auxiliary tug force (196.2kN), the mean traction in the
lifting wire increased (as previously shown in the static
analysis), but the amplification factor decreased. The same
conclusion was drawn in the experimental analysis.
2.0m wave height with a 9.0s peak period (Pierson-Moskowitz
spectrum). Table 8 presents the results of the simulations, which took into account the lifting wire, tether tractions and the
geometrical configuration of the systems (distance and angle
of BTA and lifting wire, as defined in Fig. 12). It could be seen
that for all cases, there was no risk of collision between the
BTA and the lifting wire. The minimum distance between BTA
and lifting wire was 17.5m for all cases.
Figures 37, 38, 39 and 40 present the dynamic amplification
factor and the maximum traction in the lifting wire and in the
tether. Several operational problems could be identified from
these results:
• Large forces in the tether that reached 1800kN with an
amplification factor of up to 6.5;
• Lifting wire and tether slackening for almost all cases;
• An auxiliary tug force of 196.2kN that reduced the amplification factor in the lifting wire from 3 to 2.4, but increased
the tether amplification factor (from 6.0 to 6.5).
The time series of the lifting wire traction for the case
BEPA450830A is shown in Fig. 36. It corresponds to the
case with regular wave of 2.25m amplitude and 11.78s
period, lifting wire length of 60m and Faux=98kN. For this
numerical simulation case, the maximum value was taken
as the mean of the peak values. This value was chosen since
a large variability was observed, due to numerical convergence problems in the integration of the cable numerical
model. The mean (dashed line) and maximum (continuous
line) values of the traction are also indicated on the plots.
It must be stressed that, although the dynamic amplification
factors were very similar for the experimental and numerical results, the absolute values of the mean and maximum
tractions were not very close. Such discrepancies may be
explained by differences in the static configuration of the
experiments. A visual (and rough) procedure was used in
the initial experimental set-up, and the static configurations presented in Fig. 12 and 14 could not be reproduced
accurately.
Table 7 Lifting Wire Traction - Numerical and Experimental
Results - Regular waves - Case 2.
Irregular waves - real system behavior
The behavior of the systems under real sea conditions
was predicted by means of numerical simulations. The
simulations considered several lifting wire lengths, and a
86
Fig. 35
Comparison of the dynamic amplification factor between experimental
and numerical simulation results - Regular Waves (Case 2).
Fig. 36
Lifting wire traction - Length = 60m; Amplitude = 2.25m;
Period = 8.80s; Faux = 98kN (BEPA 45 08 30A).
Fig. 38
Fig. 37
Irreg. waves; Lifting wire length 60m; Faux=98kN
Fig. 39 - Irreg. waves; Lifting wire length 100m; Faux=98kN
Irreg. waves; Lifting wire length 60m; Faux=196.2kN
87
6
New launching procedure
The experimental and numerical analysis showed that one of the
launch steps (the launch of the anchor connected to the MWA) in
the operation is critical for the line behavior. Since the traction
in the cables may reach large values, slackening may occur. A
high dynamic amplification was attained because the negative
submerged weight of the MWA reduced the mean traction of
the lifting wire. Therefore, the operation was considered unsafe
by the engineers and by the operational staff.
The numerical simulations indicated that the slackening may
occurs for 2m significative wave height. Smaller heights were
not verified, so a precise weather window cannot be obtained
from the analysis. However, only for an illustrative purpose, if
one considers that 2m is the limiting environmental condition,
such weather limitation is very restrictive considering Campos
Basin scenario. In that basin, in more than 55% of the time the
significant wave height is larger than 2m. During the winter,
this occurrence increases to approximately 74%.
A novel procedure was proposed to prevent the launch of the
MWA connected to the anchor in Case 2: launching procedure
for the anchor should follow that presented in Case 1. In this
case, the anchor would be launched towards the sea-floor, and a
ROV would be used to adjust the fine positioning of the anchor
on the sea-floor. A cable connecting the anchor to an auxiliary
vessel would be used in this step to assist the anchor positioning.
Afterwards, the MWA would be launched alone using a provisory heavy chain connected to it to increase its submerged
weight. The auxiliary vessel would also be used here. Thus,
the MWA (and the heavy chain) would descend to the anchor,
and a ROV would then connect the tethers and discard the
heavy chains.
Fig. 40
Irreg. waves; Lifting wire length 100m; Faux=196.2kN
Table 8 Lifting Wire and Tether Traction - Numerical and
Experimental Results - Irregular waves - Case 2.
With this novel procedure, the launch of the MWA and heavy
chains should demonstrate similar dynamics to that of the anchor
launch in Case 1. Table 9 presents a qualitative comparison
between the procedure studied in the present paper (Anchor
connected to the BTA) and the new procedure (two-stages
launching). In fact, the major advantage of the new procedure
is related to the overall dynamics of the systems, that reduces
the occurrence of cable slackening and the probability of line
rupture during installation.
Table 9 Comparison between procedures for MWA
installation.
88
7
Conclusions
The present paper addresses a methodology to analyze complex
offshore operations involving sub-sea installation and several
support vessels. Simple experiments were used to validate a
numerical simulator that was then used for complex simulations
of the complete installation operation under real environmental
conditions.
A procedure has been proposed for the installation of a Mid
Water Arch – MWA, which consists of a structure to provide
riser support. The installation would involve two vessels and
several cables connecting them to the MWA components.
The analysis presented showed that during one of the steps of
the launch operation (the launch of the anchor connected to
the MWA), the traction in the cables may reach large values,
and slackening may occur.
A novel launch procedure has been proposed and successfully applied to the installation of more than 3 MWAs in the
Campos Basin.
FERREIRA, M.D. (2002), “Coupled hydrodynamic analysis
of an AHTS and a box structure in waves”, Proceeding
of International Offshore and Polar Engineering Conference, ISOPE.
LIMA, J.M.T.G., Kuppens, M.L., Silveira, P.F., Stock, P.F.K.
(2008), “Development of subsea facilities in the Roncador Field (P-52)”, Offshore Technology Conference,
OTC, Houston, TX, USA.
SANTOS, M., Neves, C., Sanches, C. (2009), “Y-Method for
subsea equipment installation”, DOT Deepwater Offshore Technology Conference, Houston, TX, USA.
NISHIMOTO, K., Ferreira, M., Martins, M., Masetti, I., Martins
Filho, P., Russo, A., Caldo, J., and Silveira, S. (2003),
“Numerical offshore tank: development of numerical
offshore tank for ultra deep water oil production systems”. In Proceedings of the International Conference on
Offshore Mechanics and Arctic Engineering - OMAE,
Cancun, Mexico.
OCIMF, (1994), “Predictions of wind and current loads on
VLCCs”, Oil Companies International Marine Forum.
PINKSTER, J.A., (1988), “Low frequency second order wave
exciting forces on floating structures”, PhD Thesis, Delft
University of Technology, The Netherlands.
Acknowledgements
The authors gratefully acknowledge Petrobras for supporting
the research project conducted at the University of São Paulo.
The second author acknowledges CNPq, the Brazilian National
Research Council, Research Grant 301686/2007-6.
ROWE, J.S., Mackenzie, B., Snell, R. (2001), “Deepwater
installation of subsea hardware”, Proceedings of the
10th Offshore Symposium, SNAME, Houston, TX, USA.
S IMOS , A.N. ; Tannuri, E. A. ; Pesce, C.P. ; Aranha, J.
A. P.(2001), “A quasi-explicit hydrodynamic model
for the dynamic analysis of a moored FPSO under
current action”, Journal of Ship Research, v. 45, n.
4, p. 289-301.
TANNURI, E.A., Morishita, H.M. (2006), “Experimental and
numerical evaluation of a typical dynamic positioning
system”, Applied Ocean Research, vol. 28 pp. 133-146.
References
ARANHA, J.A.P. (1994), “A formula for wave damping in the
drift of a floating body”, Journal of Fluid Mechanics,
vol. 272, pp.147-155.
FALTINSEN, O.M. (1990), “Sea loads on ships and offshore
structures”, Cambridge, Cambridge University Press,
England.
FERNANDES, A.C., Santos, M., Barreira, R., Ribeiro, M. (2006),
“Pendulous installation method prospective model testing
and numerical analysis”, Proceedings of the International
Conference on Offshore Mechanics and Arctic Engineering - OMAE, Hamburg, Germany.
89
Appendix - TPN Description
The TPN is a time domain numerical procedure designed for
the analysis of moored and DP offshore systems. The inputs
of the simulator are:
• Floating body main parameters (dimensions, mass matrix,
etc.);
• Aerodynamic drag coefficients (following standard given
in OCIMF, 1994 );
• Current coefficients (following standard given in OCIMF,
1994) or hydrodynamic derivatives;
thrust allocation, must be used to distribute control forces
among thrusters. It guarantees minimum power consumption
to generate the required total forces and moment, positioning
the vessel. At last, a control algorithm uses the filtered motion
measurements to calculate such required forces and moment.
Normally, a wind feedforward control is also included, enabling
to estimate wind load action on the vessel (based on wind sensor
measurements) and to compensate it by means of propellers.
Furthermore, the simulator also includes models for propellers,
taking into account their characteristics curves, being able to
estimate real power consumption and delivered thrust. It also
evaluates time delay between command and propeller response,
caused by axis inertia.
• Hydrodynamic coefficients (potential damping, added
mass, first and second order wave force coefficients);
• Environmental conditions (wave and wind spectra, current);
• Mooring and risers system characteristics;
• Thrusters characteristics and layout;
• DP modes and parameters.
The non-linear time-domain simulation runs in a parallel
cluster computing system and outputs time series describing
the motions of up to two floating unities (FU) in six degrees
of freedom (6DOF), tensions on the mooring lines and hawser,
propellers thrust and power, etc., and a corresponding statistical
summary. 3D visualization outputs are also available.
The floating body high frequency motion (HF) due to the wave
action can be evaluated in two different ways. In the simpler
one the HF motion evaluated by the RAO is added to the low
frequency motion (LF) that is calculated by the 3rd order
Runge-Kutta integration method. Alternatively, the wave 1st
order forces are applied to the body and all motion components
are obtained dynamically solving the equations of motion. The
current force can be evaluated through 3 different models:
OCIMF Model, Cross flow Model, Maneuvering Model or
Short Wing Model (Simos et al, 2001). It is possible to analyze
3D constant or oscillatory current profile. The simulator allows
constant wind and gusty wind. The wind spectra implemented
in the code are Harris, Wills and API. The wave can be regular
and irregular. For irregular waves the spectra formulations available are Pierson-Moskowitz, JONSWAP and Gaussian. The
wave first and second-order effects are modeled (see Faltinsen,
1990 and Pinkster, 1988) and wave-drift damping effects are
included according to Aranha, 1994. The wave coefficients are
evaluated by WAMIT (Wamit, 2000).
Three main classes of algorithms used in commercial DP
systems are also implemented in TPN (Tannuri and Morishita, 2006). A low-pass filter, called wave-filter, is employed
to separate high-frequency components (excited by waves)
from measured signals. Such decomposition must be performed because the DP system must only control low-frequency
motion, since high-frequency motion would require enormous
power to be attenuated and could cause extra tear and wear
in propellers. Furthermore, an optimization algorithm, called
90
Model test philosophy for FPSO’s in deep brazilian waters
Mamoun Naciri
Single Buoy Moorings Inc.
[email protected]
Abstract
Recent discoveries in the deep waters of the Santos Basin Offshore Brazil are paving the way for field developments based on spread
or turret moored FPSO’s. For decades, model testing of moored floating structures has been a corner stone during project execution.
There were mainly two reasons 1) to obtain design loads to feed into the structural design and 2) to verify with site specific meteocean conditions that no unforeseen phenomena were taking place. The advent during the last two decades of increasingly efficient
analysis tools and computer hardware has made the first reason for performing model tests less pressing. There remain, however, a
caveat that analyses results could be proven to be in good agreement with experiments.
Invitations To Tender (ITT) received by SBM in recent years for FPSO’s to be deployed in deep water Offshore Brazil have
systematically called for model tests and for calibration of the numerical analysis tools used for design against project specific
experimental results.
The paper will present SBM thoughts on optimization of the model test Scope of Work in the larger context of the FPSO engineering
and design phase. An example from a recently executed FPSO project offshore Brazil will be selected for illustration.
Keywords
Brazil, FPSO, Model tests, Deep water, Sea keeping, Station keeping
1
Introduction
It is anticipated that the recent discoveries in the Santos Basin pre-salt layers in the Tupi, Iara and Guara fields will nearly double
Brazil’s current oil reserves (see Upstream 2009). Development of these and other fields in this basin will require a large number
of FPSOs moored in water depths in excess of 2000m.
Recent Invitations To Tender (ITT) received by SBM for FPSO’s to be deployed in deep water offshore Brazil almost invariably
stipulate that model tests be performed. In some instances, the nature of tests to be performed is clearly spelled out e.g. wind tunnel
tests, sea-keeping tests and station-keeping tests.
With model tests so much part of the Offshore FPS culture for decades, it is understandable that the abovementioned requirements
be found in ITT’s. Nevertheless, it is worthwhile to step back and ponder on the deep-rooted reasons why we perform model tests
and analyze in §2 if these reasons are today as pressing as they used to be decades ago. In §3 an inventory of deep water FPSO’s
recently executed by SBM indicates whether or not model tests were performed. In §4, a typical model test scope of work is presented
outlining the specific goals each category of tests is aiming to fulfill. The degree of maturity of station keeping numerical tools is
assessed in §5 through comparisons between experimental and calculated results for a recent deepwater project. Conclusions are
drawn in §6 regarding how best to combine experiments and numerical analysis.
Submitted to MS&OT on Aug 22 2010. Revised version submitted on Dec 02 2010. Accepted on Dec 18 2010. Editor: Marcelo A. S. Neves.
91
Mamoun Naciri
2
Model test objectives
For decades, model tests have been an integral part of the design of floating offshore systems. For the sake of terminology,
model tests refer in this paper to tests performed on a moored
floating structure at a reduced scale (typically 1:40 to 1:80) in
a wave basin. This facility has generally the ability to generate
wind and current as well.
The main reasons why model tests are performed have changed
over the years. One of the enduring reasons for performing
model tests is to have a reality check of a novel concept or
component thereof.
Another historical, and perhaps less enduring, reason for
performing model tests is to provide direct load/motion input
to the structural / mechanical design teams. This was particularly true in the seventies and eighties when both computer
hardware and numerical tools could not cope with analysis
requirements. This remains true for some very specific issues
concerning mature floating concepts (greenwater & slamming
for FPSO’s or slamming, air gap/run-up issues for TLP’s, VIM
for SPARs / TLP’s etc..) where analyses methods have not
yet matured to a sufficient degree to address these issues with
enough confidence.
With the advent of powerful computers and sophisticated
analysis tools, the focus of model tests has shifted somewhat
to a means to verify / validate the numerical analysis tools
employed to derive design loads and motions.
Finally, model tests are often required to provide the project
team the comfort (or perception thereof) that the designed
floating structure has passed a reality check with site specific
environments.
Altogether four reasons for performing model tests have been
identified (concept verification, source of design loads, validation / calibration of numerical tools, comfort on the global
performance). These reasons are now revisited in the specific
case of a deepwater FPSO.
With more than 90 FPSOs now in operation around the word,
the FPSO can hardly be considered as a novel concept. What
may however be considered is the envelope of already operating FPSO’s in terms of design parameters:
• Water depth;
Consider for instance a water depth of 1800m a scale model
of 1:60 and a wave basin depth of 10m. The modeled water
depth is 600m i.e. one third of the full scale depth. Mooring and
riser must be truncated. When designing a truncated mooring
system, attempt is made to conserve the following parameters:
• Horizontal mooring stiffness over a large enough excursion
range;
• Suspended weight over a large enough excursion range;
• Low frequency mooring line damping;
• Dynamic top tension.
It is not always possible to match well all parameters and
therefore a compromise should be sought depending on the
design parameters of interest. If excursions are the prominent
issue, particular attention will be paid to meeting the horizontal
mooring stiffness. If turret loads are the main concern, also the
total vertical weight should be modeled accurately. If tensions
are the prime test results, the dynamics of the truncated line
should be as close as possible to the full depth mooring line.
Figure 1 below shows a comparison of load excursion curves
for the full depth (1800m) and truncated (600m) mooring
system. A good agreement is found up to 9% of the full scale
water depth.
Recent Invitations To Tender (ITT) received by SBM for
FPSO’s to be deployed in deep water offshore Brazil almost
invariably stipulate that model tests be performed. In some instances, the nature of tests to be performed is clearly spelled out
e.g. wind tunnel tests, sea-keeping tests and station-keeping tests.
With model tests so much part of the Offshore FPS culture
for decades, it is understandable that the abovementioned
requirements be found in ITT’s. Nevertheless, it is worthwhile
to step back and ponder on the deep-rooted reasons why we
perform model tests and analyze in §2 if these reasons are
today as pressing as they used to be decades ago. In §3 an
inventory of deep water FPSO’s recently executed by SBM
indicates whether or not model tests were performed. In §4,
a typical model test scope of work is presented outlining the
specific goals each category of tests is aiming to fulfill. The
degree of maturity of station keeping numerical tools is assessed in §5 through comparisons between experimental and
calculated results for a recent deepwater project. Conclusions
are drawn in §6 regarding how best to combine experiments
and numerical analysis.
• Environments (significant wave height, wind & current
speeds etc.);
• Vessel size;
• Mooring system type;
• Mooring force;
• Suspended weight (fluid transfer system);
If one (several) of the above parameters is (are) significantly
outside the envelope of the existing FPSO fleet, one may ponder
and consider model testing. This is further discussed in §3.
Could one rely today entirely on model test results for design
verification and input to structural and mechanical design?
92
Fig. 1
Full depth and truncated load-excursion characteristics
Mamoun Naciri
Likewise, when modeling the truncated riser system, from
a mooring system design point of view, attempt is made to
conserve the following parameters:
• Suspended weight over a large enough excursion range;
• Mean current load on riser system;
Furthermore, the increased complexity of the meteocean conditions seen in recent years results in ever increasing numbers
of combinations to consider during design verification. This
clearly precludes the sole reliance on model testing. A good
example is the recent introduction of (Hs,Tp) iso-probability
contours (see Figure 3) per direction in Petrobras meteocean
condition reports instead of the single (Hs,Tp) pair seen before.
• Low frequency riser damping.
Since riser top angles are relatively small, the associated horizontal stiffness is generally small compared to the horizontal
mooring stiffness. If the number of risers exceeds the number
of mooring lines by far and the water depth is large enough
this may no longer be true.
From a mooring system point of view, the current load on risers
and the associated damping are important. Figure 2 illustrates a
typical 10-year return period current profile with the truncated
water depth indicated in red. Ten-year return period currents
are generally associated with the 100-year wind sea and wind
conditions that are critical for the design.
Fig. 3
100-year (Hs,Tp) contour for a location offshore Brazil.
In summary, it appears clearly that mooring/turret systems
design for deepwater FPSO’s cannot rely solely on model tests.
The third reason invoked for performing model testing is to
provide material for comparison, tuning and validation of the
numerical tools ultimately used in the detailed design.
3
Fig. 2
10-year current profile offshore Brazil.
The graph shows clearly that the current profile is not restricted to the model depth and that drag loads on two third
of the water column will be missed in the model basin. Furthermore, current profiles are not necessarily unidirectional
throughout the water column. In model basins, the current
can flow only in one direction and there may be limitations
in terms of the maximum current speed achieved at a given
depth for a given scale.
Experience executing deepwater
FPSO’s
SBM experience executing deep water FPSO’s worldwide
is summarized in Table A1 (Appendix A). The threshold of
deepwater has been set to 500m. The main parameters listed
in the abovementioned table are:
• Country;
• Client;
• Water depth;
In summary, it will be difficult to recover the full depth current induced mean loads with the truncated riser system – and
likewise for low frequency riser damping. This means that
design excursions cannot be obtained reliably from a model
test when such drastic truncation is required.
• Vessel deadweight capacity;
Furthermore, as the low frequency response of moored structures is affected by wave groups, a single realization of a sea
state is clearly not enough to obtain statistically reliable results.
If design parameters were to be derived from tests, numerous
realizations of a given sea state should be used thus increasing
the test program significantly.
• Date of first oil.
• 100-year significant wave height (as a measure of the
severity of the environment);
• Number of risers/umbilicals;
For each FPSO, the table indicates whether model tests were
performed or not. Figure 4 hereafter shows the chronological
evolution of water depth. Projects for which model tests were performed are identified with an icon showing a vessel scale model.
93
Mamoun Naciri
Fig. 4
Time line of FPSO projects versus water depth.
Fig. 6
FPSO II moored in 1260m water depth in Brazil (a record
depth at the time) was not model tested by SBM. The first
model test campaign is in a much smaller water depth of
850m in the Adriatic Sea for a different Client (AGIP). From
2000 to 2008 several FPSO’s in West of Africa (ExxonMobil)
and offshore Brazil moored in less than 1400m water depth
have been executed without model testing. The Espirito Santo
FPSO in 1780m water depth was clearly outside the envelope
of previously executed FPSO’s. Furthermore, it was important
for Shell to have a good understanding of extreme roll motions
to be able to design the flex joints of the novel steel catenary
lazy wave risers (see Lavagna et al. 2009). The other two model
tests are for water depths within the envelope.
It is worthwhile to emphasize that with taut & semi-taut
chain-polyester-chain systems, the light weight of the
mid-water segment helps cope with water depth increases
without significant changes in suspended weight nor in
global performance.
Figure 5 hereafter shows the chronological evolution of 100year Hs.
The P57 FPSO distinguishes itself by being spread moored (all
other FPSO’s are turret moored) and by the number of risers
connected. For this unit water depth and Hs are well within
the envelope of past projects.
After a long spell of deepwater projects executed on both
sides of the Atlantic without any model tests, a new trend has
emerged recently favoring model test campaigns for deep
water FPSO’s offshore Brazil. The main drivers for this trend
appear to be:
• First FPSO for new operator Offshore Brazil;
• Stringent roll performance criteria in view of novel
riser system;
• FPSO outside envelope of existing units for historical
operator Petrobras i.e.
- Spread moored (most existing FPSO’s are turret moored);
- Large number of risers suspended from the side.
The above is in line with the four reasons highlighted in
Chapter 2 for performing model tests (concept verification,
source of design loads, validation / calibration of numerical
tools, comfort on the global performance). In the next chapter,
a typical scope of work for model testing a deepwater FPSO
is described based on recent ITTs.
4
Fig. 5
Time line of FPSO projects versus Hs.
The 100-year Hs value for the three most recent projects is
within the range of previously executed projects in Brazil. The
Frade FPSO neither breaks a depth record nor an Hs record
(Hs=7m) but is Chevron’s first FPSO offshore Brazil.
Figure 6 hereafter illustrates the chronological evolution of
the number of risers.
94
Time line of FPSO projects versus number of risers.
Model test scope of work
A typical model test scope for an FPSO includes wind tunnel
tests, sea-keeping test and station keeping tests. Wind tunnel
tests provide key information which will impact vessel mean
forces and headings and thus 1st and 2nd order responses.
Furthermore, as these tests are very inexpensive, they should be
mandatory for any FPSO project especially when the underwater part of the hull departs from that of VLCC (large bilge keels,
sponsons, riser balcony etc…) and/or when process equipment
on deck contributes significantly to the overall windage areas.
Sea-keeping tests are generally aimed at documenting the first
order motion and acceleration responses, freeboard exceedance
(greenwater), bow slamming (if any) in a number of wave
conditions. A typical test program will include:
Mamoun Naciri
• Heeling tests to document metacentric heights;
• Added mass, radiation damping and total wave force
RAO’s from diffraction database;
• Free floating decay tests to document natural periods in
heave, roll and pitch and associated damping;
• Convolution used to account correctly for radiation effects;
• Regular and irregular waves and/or white noise tests at
relevant incidences.
• Current load on mooring & risers and associated damping
accounted for by use of full line dynamics;
For tests in waves, the hull is restrained by an equivalent horizontal mooring system. The focus of these tests is generally on
roll motions. Effectiveness of bilge keels, impact of sponsons
or of a riser balcony on the roll performance can be assessed.
Strict roll performance criteria specified by the Client can be
verified during such sea-keeping tests. In view of the sensitivity of some process equipment (e.g. fractionation towers) to
transverse accelerations, it is highly recommended to perform
sea-keeping tests for non-conventional hull shapes.
• The turret is modeled as an independent structure to
which mooring lines and risers are attached. The FPSO
is free to weathervane about the turret vertical axis. The
numerical model is shown in Figure 7;
Station-keeping tests are aimed at verifying the global
performance of the mooring system (excursions, mooring
line tensions), of the turret system (chain table and bearing
loads) in a limited number of combinations of environments
and FPSO loading conditions. A typical test program will
include:
• Static tests to document stiffness of mooring system
(and riser system if relevant);
• Decay tests in calm water and in current for horizontal
motions with mooring and risers connected;
• Wind and current coefficients as per wind tunnel tests;
• Wave elevation imported from wave calibration test (without
FPSO);
• Wind speed is constant in the time domain simulations;
• Current speed is constant in the time domain simulations.
The depth profile measured during the calibration phase
of the model tests is implemented;
• The numerical roll damping model consists of both linear
and quadratic damping coefficients. These coefficients are
derived such that the numerical simulation of the free
floating decay matches in the time domain the measured
roll decay; preferably for up to 20 periods and with due
attention to amplitudes and phases.
• Irregular wave tests including wind and current. It is
good practice to include a few tests in waves only
(this is very helpful for the calibration of numerical tools
as discussed further in §5).
5
Effectiveness of station keeping
numerical tools
The purpose of this section is to demonstrate the maturity of
numerical tools used in SBM Offshore by comparing model test results with numerical simulations for a deep water
turret-moored FPSO in Brazilian environment. The FPSO is
converted from a VLCC hull. The mooring system consists of
three bundles of three lines each. The line composition includes
a suction pile, a bottom chain segment, a mid-water polyester
segment and a top chain segment. For simplicity risers have
been lumped into a reduced number of equivalent risers.
The numerical tool utilized for this comparison is the AQWA
suite developed and maintained by Century Dynamics Limited;
an ANSYS company.
5.1
Methodology
5.2
Illustration of the numerical model.
Test matrix
Four model tests have been selected for comparison with the
numerical model. These correspond to the oblique environment
described below:
• Waves: Hs = 7.8m, Tp = 15.4s, γ = 1.7. The direction of
wave propagation is 225°;
• Wind: Uw=34.3m/s. The wind speed is constant. The
direction the wind is blowing to is 195°;
• Current: Uc=1.02m/s. The direction towards which the
current is flowing is 180°.
The following assumptions are made for the time domain
simulations:
Fig. 7
The test matrix is shown in Table 1 below.
95
Mamoun Naciri
Table 1
5.3
Test matrix
Table 2
Measured and calculated mean headings
Wave elevation import
The calibrated wave elevation in the presence of current
and without the FPSO is used. The RMS wave elevations as
measured in the basin and as imported in the simulation are
shown below:
•
Test: RMS=1.956m
•
Simulation: RMS=1.951m
Excellent agreement is found (-0.26%). The measured and
imported wave elevations are compared in Figure 8 below.
Good agreement is found in the time domain as well.
During model tests, the FPSO will experience yaw fluctuations
due to wind and current turbulence. In contrast, the numerical
calculations assume constant wind and current speeds.
Mean turret excursions
The mean X and Y positions of the turret chain table when the
FPSO is subjected to wind, waves and current are summarized
in Table 3 below.
Table 3
Measured and calculated mean turret excursions
A very good agreement is found for the post-installation stiffness case which is designing for excursions.
Mean anchor line tensions
The mean tensions from the model tests and simulations are
compared in Table 4 below for the middle line of each bundle.
Table 4
Fig. 8
5.4
Measured and calculated mean tensions
Wave elevation import in time domain simulation tool.
FPSO under mean environmental loads
Mean FPSO headings
The mean headings from measurements and simulations are
summarized in Table 2 below. The wind direction in the simulation has been shifted by 3.5° to match the measured heading
in wind only condition. Simulated mean headings agree well
with measurements with differences less than 1° when all
components of the environment are present.
96
Agreement between tests and calculations is within less than
3% for the post-installation mooring stiffness. Simulations
Mamoun Naciri
over-estimate by 8 to 14% the mean tensions with the storm
polyester stiffness.
Table 6
RMS motions at turret in wind and current only.
Mean mooring force
The mean horizontal mooring force due to both anchoring and
riser systems are compared in Table 5 below.
Table 5
Measured and calculated mean horizontal mooring
forces.
The above turret position fluctuations due to wind and current
show the benefit of performing a few tests in waves only. The
repeatability of waves allows a one to one comparison between a given test and a simulation. This adds more value to the
calibration exercise.
5.6
The magnitude of the mean mooring force is recovered within
less than 10%. Differences are observed in the azimuth which
may be due to the presence of localized transverse current
in the basin. Note that the simulation provided conservative
mooring force values.
5.5
FPSO vertical motions at turret location
The FPSO heave and pitch motions contribute mostly to the
heave motion at the chain table. The statistics of the total
heave motion at this location are compared in Table 7 hereafter. Excellent agreement is found in terms of RMS response.
Similar agreement is found for the pitch motion. The largest
discrepancy found is 6%.
Table 7
Total heave motion at turret location.
FPSO in wind and current only
One main difference between these model tests and the corresponding simulations is that the current and wind speed
are strictly constant only in the latter. In order to assess the
variability of wind and current and its effect on the FPSO, the
envelope of the measured turret trajectory during the wind
and current only tests has been plotted in Figure 9 hereafter.
Furthermore, good agreement is found in the time domain as
shown for heave in Figure 10.
Fig. 10
Measured and computed heave at turret – fully loaded FPSO (measurements
in blue).
Figure 11 below shows time series of the measured and computed roll motions for the ballasted FPSO. A fair agreement
is found in this near head on condition (15deg off the bow).
Fig. 9
Turret trajectory envelope in wind and current only tests.
The above excursions will clearly not be accounted for in the
numerical simulations. It is furthermore not possible to filter
out these fluctuations from the tests in wind, waves and current
since there is no repeatability in the wind and current turbulence and there is no simple way to differentiate the wind and
current-induced LF content from the wave-induced LF content.
Consequently, levels of agreement in horizontal motions should
be assessed in this context. The measured horizontal RMS
motions are shown in Table 6:
Fig. 11
5.7
Measured and computed roll motion – Ballasted FPSO (measurements in
blue).
Dynamic amplification factors
The DAF is calculated as the ratio of the maximum dynamic
tension to the maximum LF mooring line tension. The DAF
has been computed for the ballasted condition with polyester
97
Mamoun Naciri
storm stiffness. Table 8 below illustrates the DAF derived for
the three most loaded lines (lines 1, 2 and 3).
Table 8
Dynamic Amplification Factors.
Note that the computed excursions are much less than the design excursions of about 130m (intact case). Offshore Brazil,
the excursions are dominated by the following components (in
decreasing order of importance): mean environment, LF and
WF loads. Owing to the severe truncation, two thirds of the
water column is missing and the current profile in the remaining (top) third of the water column is somewhat smaller than
specified (due to current generation limitations), the current
loads are drastically underestimated and thus the excursions.
Excellent agreement is found between measured and calculated
DAF values. This proves the ability of the simulation to capture
the dynamics of mooring lines in waves and current. Polyester
systems in deep water usually have very small dynamic amplification as demonstrated above.
5.8
The measured maximum excursions are recovered numerically
within less than 5%. The above excursions would have been
larger had the environment in between two bundles.
Turret low frequency excursions
5.9
LF mooring forces
The maximum LF horizontal mooring forces have been computed from both model tests and simulations. The results are
shown in Table 11 below.
Maximum mooring forces.
Table 11
The horizontal motions at the turret location have been low
pass filtered at 0.25rad/s. RMS motions are computed for
the measured and numerical times series in Table 9 hereafter.
Table 9
LF response at turret location.
A generally good agreement is found in terms of maximum
mooring forces. Furthermore, the computations provide conservative results.
5.10
Good agreement in surge is found between measurements and
simulations for the fully loaded tests. In the ballasted case, the
surge RMS is significantly underestimated. A possible explanation is the fluctuations of wind and current which lead to
notable motions (see Table 6). Another source of disturbance
is imperfection of the second order correction in the wave
making which can result in free long waves propagating back
and forth in the X-direction. The maximum intact excursions
are summarized in Table 10 hereafter.
Table 10
Maximum measured and computed intact
excursions.
Conclusions
Comparisons of four model tests with simulations have been
carried out. The environmental condition is non-collinear with
a 100-year return period. Two loading conditions have been
considered for the FPSO and two polyester stiffness values
corresponding to storm or post-installation situations. The
following conclusions have been drawn:
a) The measured mean vessel headings in wind and current
only and, in wind, waves and current are well captured by
the numerical model;
b) The vertical RMS motions at turret location are recovered
within less than 6%, thus giving confidence that the mooring
and riser dynamic excitation is correctly modelled;
c) The maximum intact turret excursions are recovered within
less than 5%;
d) Design low frequency mooring forces are always conservatively estimated. Discrepancies range from 2.5% to
14.5% depending on the loading condition and the polyester
stiffness;
e) The inherent turbulence in wind and current contributes to
the low frequency excitation and response. This contribution is absent in the simulations. A one-to-one comparison
is therefore not possible. It is therefore recommended to
include in the test program some tests in waves only.
98
Mamoun Naciri
6
Figure 14. Two-dimensional damping results can then be extrapolated to the three-dimensional situation based on the extent of
bilge keels to obtain an estimate of the linearized roll damping.
Benefit of combined numerical
analyses & experiments
Numerical analyses can be used to enhance the value of the
tests. This is illustrated by two examples:
•
In one project, CFD analysis was used prior to model tests
to identify the range of bilge keel sizes to be tested for a
new hull shape;
•
In another project, station keeping analyses were performed
during the test campaign to cross-check the experimental
setup and the numerical model.
6.1
CFD analyses
Owing to the distinctive hull shape illustrated in Figure 12,
it was anticipated that bilge keels of a given height might
not be as efficient as for a standard VLCC hull shape. Two-dimensional CFD forced oscillations in roll were performed
for three different bilge keel heights (1.1m, 1.5m and 1.8m).
The roll harmonic oscillations have a 5° amplitude and a period
equal to the roll natural period. A snapshot of the vorticity field
is shown in Figure 13.
Fig. 14
6.2
Verification of harmonic analysis decomposition.
Station keeping analyses
Performing station-keeping analyses in parallel with model
testing is highly recommended as this provides a reality
check of the validity and accuracy of the numerical model
and consequently a check of the as-built data. The measured
wave elevation is imported in the station keeping software (see
Figure 15) allowing a one-to-one comparison between test
and simulation. The turret excursions and vertical motions are
respectively compared in Figures 16 and 17. Dynamic tensions
in a mooring line are compared favorably in Figure 18 when
importing in ORCAFLEX fairlead motions.
Fig. 12
Midship hull cross section.
Fig. 15
Wave elevations import in numerical tool.
Turret trajetory
Fig. 13
Snapshot of vorticity field (2D CFD calculation).
The computed dynamic roll moment is then processed using
harmonic analysis to obtain contributions in phase and out-of-phase with the forced motions. This provides estimates for 1m
ship section of the roll added inertia and linearized damping.
The quality of the simple linear reconstruction is illustrated in
Fig. 16
Turret horizontal excursions (measurements in blue & analysis in red).
99
Mamoun Naciri
example in Esperança et al. 2008). Greenwater issues, if
any, can be addressed during such tests.
• Station-keeping tests for spread or turret-moored deepwa
ter FPSO’s (with chain, steel or synthetic wire, chain line
composition) do not add significant value. These tests are
therefore not required.
Fig. 17
Turret heave motions (measurements in blue).
Acknowledgments
The author would like to thank Emmanuel Ory for performing
the 2D CFD calculations referenced in this paper and Renaud
Daran for his contribution to the station keeping calculations.
References
ESPERANÇA, P.T.T., Sales, J. S., Liapis, S., Matsuura, J. P. J. and
Schott, W. (2008), “An experimental investigation of roll
motions of an FPSO”. Proceedings of the 27th International
Conference on Offshore Mechanics & Arctic Engineering.
Estoril; Portugal. Paper # OMAE2008-57765.
Fig. 18
LAVAGNA, P., Martineau, E., Agussol, L., Wibner, C. and
Hoffman, J. (2009), “Fluid Transfer via Steel Risers on a
Turret Moored Deepwater FPSO” Proceedings of the Deep
Offshore Technology International Conference - Monte
Carlo; Monaco.
Dynamic mooring line tensions.
UPSTREAM Magazine dated July 17th 2009 - Pre-Salt Focus,
“Plenty to twist and shout about off Brazil”.
7
Conclusions
Analysis results presented in §5 demonstrate, through an
example, the maturity of station keeping numerical tools for
deep water turret moored FPSO’s. In §6, it is shown that numerical work can be performed ahead or during model tests
to optimize their value.
SBM’s philosophy regarding model testing of deep water
FPSOs is outlined below:
• Wind tunnel tests to document accurately the wind and
current coefficients should always be performed.
• Sea-keeping tests to characterize motions with special
em phasis on roll when the hull shape differs significantly
from known hull shapes are highly recommended (see
100
Mamoun Naciri
Appendix A
Table A.1 Deepwater FPSO performance record and wave basin model testing.
101
A
Tools for prediction of wave impact on FPSO’s and offshore platforms
C.T. Stansberg 1 and R. Baarholm 1, 2
1 MARINTEK, Trondheim, Norway
[email protected]
2 Present affiliation: Statoil, Trondheim, Norway
Abstract
Improved engineering tools and methods for the prediction of loads from wave impact on FPSO’s and offshore platforms in severe
sea states are described. Strongly nonlinear wave-structure mechanisms are involved, and a semi-empirical approach is established
based upon present state-of-the-art general knowledge within marine technology. It has been validated through systematic comparisons with model test data. Tools are developed for prediction of air-gap and wave-in-deck loads on platforms, and green water and
bow flare slamming on FPSO’s. Robust and practical methods are emphasized. The principles of the approach and examples are
presented. In spite of the relatively simple models assumed, promising comparisons to experiments are obtained.
Keywords
Offshore strutures, wave impact, prediction tools
1
Introduction
The prediction of wave impact on offshore structures is complex, especially in stormy sea states. Strongly nonlinear mechanisms
are involved, including nonlinearities in the waves themselves, as well as in the interaction with structures. Standard hydrodynamic
tools for engineering use generally do not handle this in a consistent and robust way. Semi-empirical methods have been developed during the last decade taking into account learning from experiments and combine with available models (Hellan et al., 2001;
Buchner, 2002). Fully nonlinear numerical modeling and CFD is also in strong development as shown in Kleefsman et al. (2004),
Collichio & Greco (2007) and Vestbøstad (2009), and will continue to be so in the future. Still, for design load estimation in the
industry, model testing continues to be generally recommended as an important tool, either used in combination with numerical /
analytical tools, or even used directly for load estimation.
Recently, improved practical procedures and tools for robust prediction have been established within a 2-year JIP research project,
“Wave Impact Loads JIP” (Stansberg et al., 2007 & 2010). Types of studies include FPSO’s as well as fixed and floating platforms.
State-of-the-art industry knowledge is combined with extensive model test experience and a critical evaluation of governing parameters. The purpose is to make optimal use of today’s knowledge in implementation into better prediction models, in particular for
use at an early stage of a design process. Simple software and analytical tools have been developed and validated for green water
and bow flare slamming on FPSO, and for wave impact on platform decks. In addition, practical guidance and procedures have
been established, and basic studies of slamming phenomena have also been carried out. In the present paper, a brief presentation
of the tool development is reviewed, and numerical examples are demonstrated. Governing mechanisms and validity of the models
are also commented.
2
Wave impact loads in random waves
Wave impact loads includes extreme global loads inducing such as ringing and whipping, global impact loads on fixed and floating
platform type structures, and local impact loads due to green water, bow slamming and wave-in-deck impact forces. Typical problems
are schematically illustrated in Fig. 1. Because of the strongly nonlinear effects that may occur, and limited accuracy or robustness
of available theoretical models for such applications, the loads are often estimated by use of model tests in offshore basins.
Submitted to MS&OT on Aug 18 2010. Accepted on Dec 22 2010. Editor: Marcelo A. S. Neves.
103
C.T. Stansberg and R. Baarholm
due to a local submergence, and can become much higher. On
the other hand, it has a more localized structure and much shorter duration and its total force effect depends significantly on
the averaging over an area or a time window. Thus the resulting
average pressure over a given plate area is clearly decreasing
with the integrated area, and a similar reduction applies for the
momentum integrated in time. Critical conditions occur when
the integrated impact forces over the defined area are still above
a certain threshold, and the duration of the force is long enough
to have a response effect on the structure. In some cases there
is also a dynamic coupling between the structural response and
the hydrodynamic loading, leading to hydro-elastic impacts.
Furthermore, in very high and energetic waves, global forces
from an impact can also be critical.
Fig. 1
Upper: Platform air-gap and deck impact. Lower: Green water on FPSO
Critical conditions are usually governed by high or energetic
waves, the local hull geometry and possible wave-induced
motions. Shorter wave periods are typically more critical
than longer periods with the same height, due to the increased
amplification of the wave-hull interaction, and due to relative
differences in phase and amplitudes between the floater and
the waves (for ships and floaters). In steep waves these mechanisms may be strongly influenced by nonlinear effects.
Also the relative heading between the structure and the waves
can be important.
The theoretical prediction of the resulting wave impact pressures and forces is complex. Contrary to the prediction of ship
and platform global forces and motions, linear wave-structure
interaction modelling cannot be applied for the impact description. Strongly nonlinear mechanisms are involved. Also the
time and space scale is much lower; durations are typically in
the order of 0.05s - 0.3s full scale. Various advanced theoretical
tools for this do exist or are in development, at least for parts
of the problem. Still the robustness for direct use in design is
presently limited, especially for 3D problems.
In the following description, new practical tools and procedures
for platform and FPSO problems are presented. Standard linear
tools and analyses form much of the background, but these
are here assumed to be known, and the nonlinear mechanisms
which are special to the new procedures are highlighted.
Fig. 2
Photos from model tests. Upper: GBS in extreme wave. Lower: Wave
slamming at FPSO bow
Examples from experiments with structures in extreme waves
are shown in Fig. 2.
Wave impact or slamming loads on a ship or platform occur
when a dynamic water wave surface hit a plate structure under
a low angle and at a high enough relative velocity to generate
a very rapidly increasing, local pressure field. This pressure
comes in addition to the wave-frequency “quasi-static” pressure
104
3
Platform air-gap and deck
impact loads
3.1
Relative waves, upwelling and air-gap
For fixed slender platform substructures such as jackets, the
relative waves and air-gap is found directly from the incident
wave crests. In steep waves, nonlinear crest contributions are
important both in deep and in shallow water, and for irregular
waves a second-order wave model is recommended, following
the theory and developments by Sharma & Dean (1981), Forristall (2000) and Stansberg (1998). A simplified crest height
probability distribution for deep water is:
(1)
and the corresponding expected extreme crest height is:
(2)
where AR is the expected crest height from the linear (Rayleigh) model. Here kp is the wave number corresponding to the
spectral peak wave frequency fp. The largest crest heights are
increased typically by up to about 15% - 20% in the steepest
spectra, relative to a linear Gaussian model with Rayleigh
distributed crests. Extreme wave heights are not significantly
affected by second-order mechanisms.
For large-volume column-based platforms, such as Semis,
TLP’s, Spars and GBS’s, the relative waves are amplified due to
the hull, and corrections must be made. For floating structures
the platform motions must be taken into account. From linear
wave diffraction theory, linear amplification transfer functions
are easily found from standard industry analysis tools, e.g.
WAMIT (2005), both for floating and fixed structures.
Fig. 3
Spatial increase in local contribution near column (“upwelling”): B(r)
as a function of distance from column wall.
Fig. 4
Large relative wave crests in front of fixed column in steep irregular wave;
predictions vs. model tests. Upper: Linear model. Lower: Modified linear model.
However, in steep waves nonlinear wave-hull contributions
grow important, see e.g. Nielsen (2003) and Stansberg &
Kristiansen (2006). Second-order diffraction models can often
be an improvement, at least in moderate waves, but they are
often laborious in use, and in 100-year and 10000-year sea
states it is also known that they can lead to large and unphysical over-predictions, as described in Sweetman et al. (2002)
and Teigen & Niedzwecki (2003), and corrections are needed
(Stansberg et al., 2007); Stansberg & Kristiansen, 2006). For
practical applications, a modified linear model has therefore
been established with an empirical correction procedure to take
into account nonlinear effects based upon observations from
systematic comparisons to model tests:
The largest crests are amplified as:
(3)
Here the first term Ad,0 = the linear wave elevation crest from
theoretical calculations (e.g. WAMIT); kpAmax is a rough measure of the expected extreme wave steepness, with kp being
the wave number corresponding to the spectral peak frequency
and Amax being the expected linear extreme crest height in the
incident wave. The factors B(r) . C(kpa) are defined as:
when r < b; and zero elsewhere
(4)
(r distance from column wall; b 2a for single or upwave
columns; b 0.5a for aft columns, and a column radius)
when kpa < 1; and zero otherwise
(5)
The formulation in Eqs. (3-5) implies 1) a general increase in
the whole field under the deck, of extreme diffracted crests
relative to linear estimates, plus 2) an additional increase
(“upwelling”) in a region of about 1 column radius, depending
on the location of the column. See the spatial variation of B(r)
in Fig. 3. The near-column increase only applies in the upwave
+-90 deg half-circle zone around any given column, and for
regions more than approx. 1m from the column walls. (That
is, thin run-up jets are not included). This nonlinear correction
formula is applicable only for the large crests.
105
Applications on a single fixed column case are shown and
compared to results from steep irregular-wave model tests in
Fig. 4. The amplified wave elevation is measured at 1.75m in
front of the column. Correlation plots between the predicted and
measured largest crests are presented, first with a linear model
and then with the modified model. The correction is seen to
work quite well in this case. The high nonlinear amplification
at this location (60%) is due to the proximity to the wall (ref.
Fig. 3); for larger distances it reduces to around 20% - 30%.
For floating platforms, the air-gap is also influenced by the
platform motions. The wave-frequency components and their
phase coupling to the amplified wave is usually quite well
predicted from standard linear diffraction analysis. Lowfrequency heave, roll and pitch motion is more complex and
requires more attention.
3.2
Deck impact loads
The procedure follows the approach in Baarholm (2005). Assuming incompressible and irrotational fluid flow, the global wave-in-deck problem can be described in terms of a boundary value
problem (BVP). By imposing conservation of fluid momentum,
a simple expression for the total vertical water entry force on the
deck due to slamming is found:
(6)
where A33=A33(t) is the high frequency added mass of the instantaneous wetted deck area and VR=VR(t) is the average value over
the wetted deck area of the relative impact velocity. The first term
on the right hand side of Eq. (6) is denoted as the slamming force
and the latter is an inertia force. This formula is well known and
used for a large number of water entry problems, including the
classic works by von Karman and Wagner (1932). In addition the
Froude-Krylov and hydrostatic force terms will contribute to the
total vertical load. The Wagner based method does not apply for the
water exit phase. In this phase a von Karman type approach is used.
For slender platforms, the forces can be derived directly from
the incident waves and their kinematics. An example with comparison of the Wagner based method (WBM) to experiments is
shown in Fig. 5.
In the case of large-volume hull structures, such as GBS, semi,
TLP and Spar, the hull will modify the wave elevation and kinematics, typically by amplification or upwelling mechanisms
due to interaction with columns and pontoons. In addition, for
floating platforms also the motions will influence the problem.
After the local elevation and kinematics have been determined,
in the present approach the global forces on the deck are found
by a procedure similar to that derived for slender platforms
above. What is needed particularly for the large-volume platform in this context is a procedure that takes into account the
influence from the hull on the wave field in steep sea states,
including both linear and nonlinear diffraction effects as well
as more local upwelling effects. The large-volume effects
will generate a more arbitrarily 3D shaped wetted deck area
as impact occurs, as compared to the slender platform case.
A simple and efficient wave-in-deck load software has been
established where a von Karman - momentum approach has
been implemented. The method has been described in /6,18/.
The software takes into account 3D effects and can be used
both for slender as well as large-volume hull based platforms.
In the latter case, the deformation of the wave field (elevation and kinematics) due to the hull is taken into account in
the boundary value problem, and a linear or a second-order
diffraction model can be chosen. A diffraction analysis code,
e.g. WAMIT, is needed for preparing.
The incident wave can be described in terms of regular Stokes
waves, as linear or non-linear random waves or measured
waves. The impact velocity and acceleration at the deck level
should be consistent with the order of the wave elevation. The
impacting irregular wave crest is fitted to a regular 2nd order
Stokes wave during the computation of the water impact force.
The method has been earlier verified to model tests with a
largevolume structure (GBS) in random waves, ref. Stansberg
et al. (2007) and Baarholm (2009). A plot from the former
reference is shown in Fig. 6.
Fig. 6
Fig. 5
106
Comparing numerical results by the Wagner method (WBM) and experimental
results: Wetted length and vertical force, 2D and 3D flow (from Baarholm, 2009)
Simulated vs. measured global vertical deck force on a GBS in an extreme
irregular wave event (Stansberg et al., 2007)
The maximum upwards force in the water entry phase is quite
well reproduced, while the predicted negative force in the exit
phase is somewhat larger than measured. This may be due to
higher-order effects in the irregular wave field diffracted by the
hull and deck structure.
Output data from the program includes:
• Wave time series;
• Wave-in-deck force time series;
• Tables with estimated extreme values, from series of irregugular wave runs with many realizations (see example in Fig. 7).
is used. An analytical, semi-empirical expression is used for
the maximum elevation and corresponding water velocity at
selected points on deck (xm is the distance from the bulwark)
(7)
where hi,D0 is the classical dam-break expression = (4/9)AR,i
from hydraulic dam-break theory, Stoker (1957). The additional factors represent corrections due to deviations from the
classical dam-break condition:
is a relative addition due
to the incoming velocity, Bprop (xm) is a location-dependent
reduction factor due to time-limited dynamical behavior of
the incident water level, and Bbulwark is another reduction factor
due to the bulwark.
The corresponding water velocity is obtained by adding the
incident velocity ui and the velocity obtained from the basic
hydraulics theory (Stoker, 1957), given the input height AR,i
and the height hi,max0(xm ) at the actual location:
(8)
0
Fig. 7
4
(the mark means that we use the height calculated without
bulwark, i.e. the bulwark reduces the height, but does not
affect the velocity).
Example on statistical treatment of extremes from many realisations.
Green water and bow flare
slamming on FPSO
The overall calculation approach in the present method, previously presented in Stansberg & Berget (2009), can be defined
through 4 successive tasks:
1) Relative wave elevation and kinematics at specified
bulwark locations;
2) Green water shipping and propagation on deck;
3) Slamming loads on vertical structure on deck;
4) Bow flare slamming.
The above simple formulas are compared in Fig. 8 to more
accurate 3D numerical simulations with the software “WaveLand” (Helland et al., 2001) based upon a shallow- The
described method is implemented into a software tool, where
results from a linear WAMIT model (or similar) for the ship are
needed as input. Time-domain simulations for the ship motion
and wave diffraction are computed, and events with overtopping are further analysed with the analytical formulation
above. Tables and figures with water heights and velocities at
specified locations are then also given as output, see e.g. Fig.
9. The software also computes resulting slamming loads on a
vertical wall at a given distance from the bulwark, and on the
bulwark, see the next Section.
4.2
4.1
Green water on deck
Linear ship motion and wave diffraction is combined in time-domain simulations with nonlinear irregular wave modeling,
including kinematics (horizontal free-surface velocity) in addition to elevation. A second-order incident wave model is used,
based on the formulation in Stansberg (1998) and Stansberg
et al. (2008). A linear amplification term due to the linear 3D
diffraction, plus a small higher-order empirical term, is added.
The nonlinearities in the incident wave elevation and velocities
are quite essential in the green water and bow flare slamming
prediction. The relative elevation at the bulwark is given by
the amplified wave elevation combined with the time-varying
local vertical motion. The incident velocity is set equal to the
incident water velocity. (In “shadow” zones this is set to zero).
“Events” are defined when positive relative waves occur. From
the relative motion amplitude (height) AR,i (measured from
the deck level, not from the bulwark top), and corresponding
velocity amplitude ui , the water on deck is calculated for each
event using a modified hydraulic approach. Here a 2D approach
Impact forces
In the software, slamming forces on a vertical wall on deck, in
the case of “events”, are calculated directly from the velocities
above, by the simple well-known formulation:
(9)
where ρ is the density of water (1025 kg/m3) and CS is a
slamming coefficient. Here we choose to use CS =1.5, which
is lower than values often used for slamming calculations. This
is justified by previous green water model test observations,
and qualitatively explained by the very irregular geometry
of the propagating water surface front. The pressure is to be
interpreted as an average value over a full scale area typically
2m x 2m or similar, and not as a local pressure over a very
small area.Drag forces on small or slender bodies can be found
from the calculated velocities by use of the Morison equation.
Slamming pressures averaged over areas larger than approx.
2m x 2m on the bow flare / bulwark are also calculated for the
same wave events:
(10)
107
where βi is the angle between the wave surface and the flare.
A similar formula is recommended by DNV (2007) for the
maximum of spatially average impacts during the water entry
of a wedge with an angle β relative to a water surface, except
that they use (cotg)1.1β in stead of cotg2β. Simulated bow flare
pressures for the same test run as in Fig. 9 are shown in Fig.
10. (Velocities here go in the negative direction).
The above formulation has been validated through comparison
to model test data in Stansberg & Berget (2009). The predicted
pressures were slightly higher than measured, but taking into
account the considerable expected statistical uncertainties in impacts occurring in irregular waves, the agreement was reasonable.
Acknowledgments
This work is based upon developments made within the
“MARINTEK Wave Impact Loads” Joint Industry Project
(JIP), Phase 1, 2007-2008. Participants included: ABS, Aker
Solutions, Chevron, ConocoPhillips, Det Norske Veritas AS,
MARINTEK, Offshore Innovative Solutions, LLC, Petrobras,
SEVAN, and StatoilHydro. The sponsors are gratefully acknowledged for the permission to publish this paper.
Fig. 8
Maximum water height and corresponding velocity on deck for an FPSO
water-on-deck wave event in Hs=12m, Tp=12s. New model vs.
“WaveLand” simulations (software from Hellan et al., 2001). Head sea.
Fig. 10
Simulated bow flare slam pressures for the same case as in Fig. 9.
References
BAARHOLM, R. (2005), “A simple numerical method for evaluation of water impact loads on large-volume offshore
platforms”, OMAE 2005-67097, Proc. 24th OMAE Conf.,
Halikidiki, Greece.
BAARHOLM, R. (2009), “Experimental and theoretical study
of three-dimensional effects on vertical wave-in-deck
forces”, Paper # OMAE2009-79560, Proc., OMAE 2009,
Honolulu, HI, USA.
Fig. 9
108
Water height calculated at various locations on deck, correlated vs. the
input relative height at bulwark. All overtopping events in a 3-hrs simulation,
steep sea state Hs=12m, Tp=12s. Bulwark 1.5m.
BUCHNER, B., (2002), “Green water on ship-type offshore
structures”, Dr. Thesis, Technical University of Delft, The
Netherlands.
C OLLICHIO & G RECO M., (2007), “Application of a 2D
BEM-level set domain decomposition to the green-water problem”, International Workshop on Waves and
Floating Bodies.
SWEETMAN, B., Winterstein, S.R. and Meling, T.S. (2002),
“Airgap prediction from second-order diffraction and
Stokes Theory”, International Journal of Ocean and
Polar Engineering, Vol. 12, No. 3, pp. 184-188.
DNV (2005), “Environmental conditions and environmental
loads”, DNV Recommended Practice, DNV-RP-C205,
Chapter 8 Air-Gap and Wave Slamming.
TEIGEN, P. and Niedzwecki, J.M. (2003), “Wave diffraction
effects and runup around multicolumn structure”, Proc.,
13th ISOPE Conf., Honolulu, Hawaii, USA, pp. 137-144.
FORRISTALL, G., (2000) “Wave crest distributions: observations and second-order theory”, Journal of Physical
Oceanography, Vol. 30, pp. 1931-1943.
VESTBØSTAD, T.M. (2009), “A Numerical study of wave-in-deck impact using a two-dimensional constrained
interpolation method”, Doctoral Thesis, Norwegian University of Science and Technology, Trondheim, Norway.
H ELLAN , Ø., Hoff, J.R., and Stansberg, C.T, (2001), “A
practical design tool for wave impact on bow and deck
structures”, Paper No. 1141, Proc., PRADS’01 Conference, Shanghai, China.
K LEEFSMAN , K.M.T., Fekken, G., Veldman, A. E.P. and
Iwanowski, B.,(2004) , “An improved volume-of-fluid
method for wave impact problems”, Proc.14th ISOPE,
Toulon.
W AGNER , H. (1932), “Uber Stoss-und Gleitvorgange an
der Oberflache von Flussigkeiten”, ZAMM, 12 (4),
pp.192-214.
WAMIT, Inc. (2005), WAMIT User Manual, 6.2 ed., Massachusetts, USA.
NIELSEN, F.G. (2003), “Comparative study on airgap under
floating platforms and run-up along platform columns”,
J. of Marine Structures, 16, PP. 97-134.
SHARMA, J. and Dean, R.G. (1981), “Second-order directional seas and associated wave forces”, J. Soc. of Petr.
Eng., SPE, pp. 129-140.
STANSBERG, C.T. (1998), “Non-gaussian extremes in numerically generated second-order random waves on
deep water”, Proc., Vol. III, the 8th ISOPE Conference,
Montreal, Canada, pp. 103-110.
S TANSBERG , C.T., Baarholm, R., and Zhao, R. (2007),
“Prediction of wave impact on offshore installations in
extreme storms”, Proceedings, Deep Offshore Technology (DOT) 2007, Stavanger, Norway.
STANSBERG, C.T., Baarholm, R., Berget, K., and Phadke,
A.C. (2010) ” Prediction of wave impact in extreme
weather”, Paper # OTC-20573, Proc., OTC 2010 Conference, Houston, TX, USA.
STANSBERG, C.T. and Berget, K. (2009), “Simple tool for
prediction of green water and bow flare slamming on
FPSO”, Paper No. 79489, Proc., OMAE2009, Honolulu,
HI, USA.
STANSBERG, C.T., Gudmestad, O.T., and Haver, S.K. (2008),
“Kinematics under extreme waves”, ASME Journal of
Offshore Mechanics and Arctic Engineering, Vol. 130.
STANSBERG, C.T. and Kristiansen, T. (2006), “Non-linear
scattering of steep surface waves around vertical columns”, Applied Ocean Research, Vol. 27, No. 2, pp 65-80.
STOKER, J.J. (1957), Water waves: the mathematical theory
with applications, Interscience.
109
MS&OT – Guidelines for Authors
Title of paper
First Name Surname, Organisation, Address of corresponding author (including e-mail)
Abstract
The abstract should be a brief description of the scope of the paper, not exceeding 100 words in length
Keywords: at least 3 suitable words for indexing purposes
Nomenclature
A nomenclature is required for papers using a large number of symbols, abbreviations and acronyms. Where possible, these should
be ordered alphabetically.
Symbol 1 Definition
Symbol 2 Definition etc.
E.g.:
α
ρ
λ
Angle of attack
Density of water
Wave length
1. Introduction
This is normally the first section in the main body of the text. Please note that this section and all subsequent sections and subsections
are numbered. All main headings should be typed in bold as shown below.
2. Heading
2.1 Sub-heading
Each Section may have sub-headings. Sub-headings should be numbered and typed using lower case as above.
2.2 Sub-heading
2.2 (a) Further subsidiary heading
The sub-sections can be further divided up as above. Further subsidiary headings not in bold in lower case.
3. Manuscript format conventions
3.1 Font
The fonts to be used are the same that have been used for this page, Times New Roman and Arial. The title of the paper should be in
12 point bold capitals. Authors names should be in 10 point bold, with affiliation in 10 point regular. The rest of the paper should
be in 10 point using the font style indicated in this template.
3.2 Page setup
The final manuscript should use the style used in this template. The margins are as follows for both A4 and US letter style in order
that Acrobat versions of the manuscript can be printed on either A4 or 8.5x11 inch paper:
Top
A4
8.5x11 in.
2.5cm
1.0 in
Bottom
3.5cm
0.7 in
Left
1.7cm
0.8 in
Right
3.5cm
0.7 in
110
3.3 Page numbers
Please put page numbers on manuscript at bottom center. This will be redone by the editors in the final printed version. Total height
of printed materials including page numbers equals 23.7 cm or 9.3 inches
3.4 Figures
3.4 (a) General guidelines
w All illustrations should be clearly referenced in the text
w Figures should be placed in the main body of the text.
w Text within figures should be of a size to allow legibility even if reduced.
w Figures may be in color or in black and white. If you use color graphics please check the graphic in black and white to see
if shading or hatching is needed. Wherever possible, figures will be in black and white.
w Captions for illustrations should be typed underneath each figure.
3.4 (b) Figure format
Figures should be produced electronically where possible, in .wmf, .eps, .jpg, .cdr, .gif, or .tif formats. Excel charts should be copied
and pasted to Microsoft Word. Save another copy of each individual graphic in separate file (Fig1, Fig2, etc.) in addition to the complete manuscript file. The resolution should be at least 300dpi, and preferably above 500dpi.
Please ensure that the lettering used in the artwork/illustrations does not vary too much in size. The final font size should be about
6-8 point. Make sure that the physical dimensions of your artwork match the dimension of A4 paper.
4. Submission requirements
Authors must submit their paper IN ELECTRONIC FORMAT in word processor or Acrobat format. They must be sent to one of
the Editors.
w Papers should be sent as e-mail attachments if each file is no larger than 3MB.
w If larger than 3MB, a CD should be sent to the Editors.
w Maximum length of paper 20 pages (above 20, with editors approval).
w If there is any possibility that certain symbols may not translate to an English version of MS Word or the Acrobat version may
have missing imbedded fonts, 2 hard copies of their printed manuscript should be sent to the Editors so that proper editing to
English symbols can be done.
5. Additional information for authors
5.1 Liability
Authors are responsible for obtaining security approval for publication from employers or authorities where necessary. If they so
wish, authors should include a disclaimer at the end of the paper stating that the opinions expressed are those of the author and not
those of the company or organisation that they are representing.
6. Conclusions
The main body of the text should end with the conclusions of the paper.
Acknowledgements
Brief acknowledgements may be added.
References
References should be listed alphabetically with the year of publication just after the author’s last name aand referred to in the text
as Firth (1989):
Firth, S. (1989) - “Investigation into the Physics of Free-Running Model Tests”. SNAME Transactions, pp. 169-212.
111
Calendar of Events
COPINAVAL 2011
IWWWFB 2011
26th International Workshop on Water Waves and Floating
Bodies
Athens, Greece, 17-20 April 2011
www.iwwwfb.org/Workshops/26.htm
OTC 2011
41st Offshore Technology Conference
Houston, USA, 2-5 May 2011
www.otcnet.org/2011
XXII Pan American Conference of Naval Engineering, Maritime Transportation & Ports Engineering
Buenos Aires, Argentina, 27-30 September 2011
www.copinaval.org
Rio Pipeline 2011
Conference & Exposition
Rio de Janeiro, Brazil, 20-22 September 2011
www.ibp.org.br
MTS 2011
2nd Annual Marine Tech Summit
Busan, South Korea, 23-25 September 2011
www.bitconferences.com/mts2011/default.asp
ISSW 2011
12th International Ship Stability Workshop
Washington D.C., 12-15 June 2011
www.sname.org/sname/chesapeaksection/Events/ISSW/
Default.aspx
OTC Brasil 2011
1st Offshore Technology Conference Brasil
Rio de Janeiro, Brazil, 4-6 October 2011
www.otcbrasil.org
OMAE 2011
30th International Conference on Ocean, Offshore and Arctic
Engineering
Rotterdam, Netherlands, 19-24 June 2011
www.asmeconferences.org/omae2011
COBEM 2011
21st International Congress of Mechanical Engineering
Natal, RN, Brazil, 24-28 October 2011
www.abcm.org.br/cobem2011
ISOPE 2011
21st International Offshore and Polar Engineering Conference
Maui, Hawai, USA, 19-24 June 2011
www.isope.org
ITTC 2011
26th International Towing Tank Conference
Rio de Janeiro, Brazil, 28 August - 03 September 2011
www.laboceano.coppe.ufrj.br
IFAC 2011
18 th World Congress of the International Federation of
Automatic Control
Milan, Italy, 28 August - 02 September 2011
www.ifac2011.org
IMAM 2011
XV International Conference - International Maritime Association of Mediterranean
Genoa, Italy, 13-16 September 2011
www.imam2011.it
112
SOBENA
Sociedade Brasileira de Engenharia Naval
The Sociedade Brasileira de Engenharia Naval (SOBENA) is
the Brazilian forum for exchange of theoretical and practical
knowledge amongst naval architects and marine engineers.
It was founded in the beginning of the modern phase of
Brazilian naval construction, in 1962, with the aim of bringing together engineers, technicians and other professionals
involved in activities as: shipbuilding and ship repair, design
and other engineering services, maritime transportation,
waterways, ports, specialized cargo terminals, ocean and river
transportation economics, marine environmental protection,
offshore support bases, offshore logistics, naval aspects of
offshore exploration and production, construction and
conversion of platforms and other offshore vessels.
SOBENA is a non-profit civil society, declared a federal public
utility by Decree No. 97589/89, which since its foundation is
aimed at promoting technological development in the above
activities through courses, conferences, seminars, lectures
and debates. SOBENA is a source of reference called upon
to provide its opinion on matters of public interest and has
also been politically active, expressing its views concerning
topics of national relevance related to its areas of activity.
Following the evolution of the industry in the past years, SOBENA has started to include activities related to offshore oil
exploration and production, holding events for professionals
of those areas. As a member of the Mobilizing Committee of
the National Petroleum Industry Organization (ONIP), SOBENA has been taking part in various subcommittees which
are seeking to create conditions to promote the development
of the Brazilian naval and offshore construction industry.
SOBENA has signed affiliation agreements with the Institute
of Marine Engineers (IMarEST), with headquarters in London, England and cooperative agreement with The Society
of Naval Architects and Marine Engineers (SNAME), from
the United States of America.
President
Alceu Mariano de Melo Souza
Vice-President
Floriano Carlos Martins Pires Jr.
Regional Director - Bacia de Campos
Aribel de Oliveira Lopes
Regional Director - São Paulo
Carlos Daher Padovezi
Regional Director - Amazônia
Fábio Ribeiro de A. Vasconcellos
Administrative Director
Ana Paula dos Santos Costa
Financial Director
Luiz Sérgio Ponce
Technical Director
Luis Felipe Assis
Associated Directors
Francisco Roberto Portella Deiana
Luiz Carlos de A. Barradas Filho
Anderson Mariano Carvalho
Address:
Av. Presidente Vargas, 542 - Gr. 713
Centro - CEP 20071-000
Rio de Janeiro - RJ - Brasil
Telephones: [+55](21) 2283-2482
Telefax: [+55] (21) 2223-3440
E-mail: [email protected]
Site: www.sobena.org.br
CEENO
Centro de Excelência em Engenharia Naval e Oceânica
The Centre of Excellence in Naval Architecture and Ocean Engineering (CEENO) was created in 1999 as a
result of a joint initiative of four Brazilian institutions (COPPE, IPT, PETROBRAS and USP), traditionally
involved in scientific and technological development applied to marine activities.
As a Centre of Excellence, CEENO aims to integrate facilities and human resources, developing theoretical and
experimental methods, giving strong support for consolidation, expansion and improvement of the maritime
activities in Brazil and worldwide.
CEENO has been involved in relevant projects on Offshore Engineering and Ship Design & Construction.
S O B E N A

Link - Sobena

Transcrição

Documentos relacionados

instructions to prepare a paper for the european congress on

Read more

Ionosphere Properties and Behaviors - Part 2

Solutions That Work - The Comprehensive Data Dictionary

Radial Shock Wave Devices can Generate

A TIME DOMAIN RANKINE PANEL METHOD FOR 2D

The Art of

PDF

Apresentação do PowerPoint

als PDF herunterladen

Author`s personal copy

TLJ70060020.

pdf

3.5.0 AUTO ENG

Gigantism and Acromegaly Due to Xq26 Microduplications and

DevInfo 6.0 User`s Guide

rigamarole - Diamond Offshore Drilling

Informações Candidatos 2014

Pigments based on Cr and Sb doped TiO 2 prepared by

Knowledge behind conservation status decisions: Data basis

Permanent hair removal of white, grey and light

Frontiers of the Roman Empire - The European Dimension of a