Front Cover
 Title Page
 Table of Contents
 List of Tables
 List of Figures
 List of symbols
 General concepts and previous...
 Centerwell mathematical model...
 Experimental results
 Model calibration and centerwell...
 Application and conclusions
 Biographical sketch

Title: Heave reponse of a deep draft spar platform with a centerwell
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Permanent Link: http://ufdc.ufl.edu/UF00091409/00001
 Material Information
Title: Heave reponse of a deep draft spar platform with a centerwell
Series Title: Heave reponse of a deep draft spar platform with a centerwell
Physical Description: Book
Language: English
Creator: Miller, William
Publisher: Coastal & Oceanographic Engineering Dept. of Civil & Coastal Engineering, University of Florida
Place of Publication: Gainesville, Fla.
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Bibliographic ID: UF00091409
Volume ID: VID00001
Source Institution: University of Florida
Holding Location: University of Florida
Rights Management: All rights reserved by the source institution and holding location.

Table of Contents
    Front Cover
        Front Cover
    Title Page
        Page i
        Page ii
    Table of Contents
        Page iii
        Page iv
    List of Tables
        Page v
    List of Figures
        Page vi
        Page vii
        Page viii
    List of symbols
        Page ix
        Page x
        Page xi
        Page xii
        Page 1
        Page 2
        Page 3
        Page 4
        Page 5
        Page 6
    General concepts and previous work
        Page 7
        Page 8
        Page 9
        Page 10
        Page 11
        Page 12
        Page 13
        Page 14
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        Page 25
        Page 26
    Centerwell mathematical model development
        Page 27
        Page 28
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        Page 44
        Page 45
        Page 46
    Experimental results
        Page 47
        Page 48
        Page 49
        Page 50
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        Page 63
        Page 64
        Page 65
        Page 66
        Page 67
    Model calibration and centerwell heave dynamics
        Page 68
        Page 69
        Page 70
        Page 71
        Page 72
        Page 73
        Page 74
        Page 75
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        Page 84
        Page 85
        Page 86
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        Page 88
        Page 89
        Page 90
    Application and conclusions
        Page 91
        Page 92
        Page 93
        Page 94
        Page 95
        Page 96
        Page 97
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        Page 129
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        Page 131
        Page 132
        Page 133
        Page 134
        Page 135
    Biographical sketch
        Page 136
Full Text




William Miller, Jr.










I would like to thank Dr. Sheppard, my advisor and supervisory committee

chairman, for his help and support during my research and Dr. Ahmed Omar for helping

me to understand some of the processes involved. Thanks also go to the staff of the

Coastal and Oceanographic Engineering Laboratory. Special thanks go to Chuck

Broward who designed and built the instrumentation used in my experiments and Vernon

Sparkman who built the model. None of this would be possible without their help. My

gratitude also goes to Dr. Sheng for his assistance in helping me understand the intricacies

of numerical solutions.

Many thanks go to my classmates and friends in the Coastal Engineering

Department at the University of Florida, especially Becky, Al, Eric, Greg, Mike, Hugo,

Santi, Mark, Chris and Wayne, all of whom acted as sounding boards and lent

encouragement when necessary. Special thanks go to Dr. Dean and Dr. Thieke for their

patience and understanding in scheduling my defense.

Finally, I'd like to thank my parents, Bill and Claudette, who have always

supported me in my endeavors.

Partial funding for this study was provided by the Offshore Technology Research

Center (OTRC) of Texas A&M University.



A CK N OW LED G M EN TS .............................................................................................. ii

LIST OF TA BLE S .............................. ... .........................................................v

LIST OF FIGURES................................... ... ...............................................vi

LIST O F SY M B O L S......................... .. ...................... ........................ ....................... ... ix

A B STRA CT .................................................................... ........................................ xii

CHAPTER 1: INTODUCTION............................. ............................. 1

CHAPTER 2: GENERAL CONCEPTS AND PREVIOUS WORK.................................7

Spar Buoy/Platform M otion...................................................... .. ........................7
Effects of a Moonpool or Centerwell ..................................... ..................... 21
Large Spar Platform s ........................ ......... ................... ............. ......................... 25


Momentum Equation for Heave............................................................................27
Inviscid Forces on the Fluid in the Control Volume................................................31
V iscous D am ping............................... ........... ..... ... ..................... .................36
Complete Equation for Heave Motion...............................................................36
Analytic Approximation to the Complete Equation for Heave ................................. 40

CHAPTER 4: EXPERIMENTAL RESULTS................................................................47

Model Description and Scaling.......................................................47
Free R response T ests ..................................................................................... ... 49
Configuration for the Forced Motion Experiments ............................................... 52
Experimental Results ...................................................... 56


D Y N A M IC S ................................................................... .................... 68

M odel Calibration............................... .. .... .. ....................... 68
Sensitivity of the Centerwell Heave Response Dynamics to Secondary Effects ........ 80
Summary of Centerwell Heave M otion................................................................ 86

CHAPTER 6: APPLICATION AND CONCLUSIONS ............................................... 91

Application of Centerwell Heave Dynamics to the Spar Platform............................. 91
Affect of a Centerwell on the Spar Platform ........................................................... 94
Conclusions and Questions for Further Study..................................... 103


Solution to the Centerwell Equation of M otion.................................................... 105
Coupled Spar and Centerwell Equations of Motion.............................................. 112

APPENDIX B: FORTRAN AND MATLAB PROGRAMS ....................................... 114

Fortran 90 Numeric Solution to Centerwell Governing Equations ....................... 114
Fortran 90 Analytic Solution to Centerwell Governing Equations........................ 120
MATLAB Analytic Solution to Centerwell Governing Equations.......................... 122
MATLAB Analytic Solution to Coupled Spar and Centerwell Governing
Equations ............................................... 125

LIST OF REFEREN CES............................................................. .................... 131

BIOGRAPHICAL SKETCH .............................................................. 136


Table page

Table 3.1, Comparison of term in Aalbers' equation (Aalbers, 1984) and the
param eters in Figure 3.1........................................................................... 38

Table 3.2, Order of Magnitude of Individual Terms in Heave Equation.......................... 43

Table 3.3, Orders of Magnitude of Complete Terms .................................................. 44

Table 4.1, Orifice Plate D iam eters............................................................................ 49

Table 4.2, Summary of Free Response Damped Natural Frequency (fd),......................... 51

Table 4.3, Estimated experimental error values. ............................................................61

Table 4.4, Synchronous motion frequencies (on) by orifice size. ................................... 63

Table 5.1, Added mass coefficient derived from the .................................................... 69

Table 5.2, Maximum amplitude ratio comparison for non-linear damping only ..............77

Table 5.3, Maximum amplitude ratio comparison for combined linear and non-linear
dam ping .............................................................. ............................. . . 80

Table 6.1, Important dimensions for coupled equation solutions........................................ 94


Figure page

Fig. 1.1, Drilling and Production Platforms (Clauss et al, 1992)..............................1

Fig. 1.2, Artist's impression of a Spar platform in the Gulf of Mexico
(Deep Oil Technology/ Rauma Repola)..........................................................

Fig. 1.3, Chevron Spar Platform Mooring Arrangement (Glanville, 1991) ......................

Fig. 1.4, Chevron Spar Platform Elevation View (Glanville, 1991).................................

Fig. 2.1, Spar Buoy Arrangement (Lewis, 1967, p. 628) .............................................. 8

Fig. 2.2, Added mass coefficients for heave motion of a vertical cylinder, G = gravity,
A = Diameter, T = draft (Sabuncu and Calisal, 1981, p. 41) ......................... 18

Fig. 2.3, Added masses of vertically oscillating components of offshore structures
(Clauss et al, 1992, p. 281) ............................................... 19

Fig. 2.4, Composite Spar Platform.................. ......................................................20

Fig. 2.5, Examples of Semi-Submersible Platforms (Clauss, 1992) ..............................21

Fig. 3.1, Spar and Centerwell Configuration and Definition of Terms. (not to scale)....... 28

Fig. 3.2, Control Volume and Control Surface for deriving the Centerwell Equation
of Motion. (not to scale) ........................................................................... 28

Fig. 3.3, Spar diagram to scale .................................................................... ................... 33

Fig. 4.1, Scale drawing of model and orifice plates.................................................... 50

Fig. 4.2, Fram e and drive assembly. ............................................... ........................ 55

Fig. 4.3, Plot of synchronous frequency with orifice diameter ratio................................. 63

Fig. 4.4, Results for experiments conducted with an orifice area ratio of 0.52................ 64

Fig. 4.5, Results for experiments conducted with an orifice area ratio of 0.64................. 65

Fig. 4.6, Results for experiments conducted with an orifice area ratio of 0.74................. 66

Fig. 4.7, Results for orifice experiments conducted with an orifice area ratio of 1.0
(i.e. no orifice). .................... ................................... 67

Fig. 5.1, Added mass coefficient vs. the cube of the orifice diameter ratio .................... 70

Fig. 5.2, Mathematical model solutions for area ratio 0.74, non-linear damping only
w ith pis = m2 = 0..................... ...................................................... 75

Fig. 5.3, Numerically determined mean centerwell level variation with frequency............ 76

Fig. 5.4, Amplitude ratio vs. frequency plot near the synchronous frequency for
orifice area ratio 0.74 .................. ........... ....................... .......................... 78

Fig. 5.5, Linear damping coefficient vs. orifice area ratio.............................................. 78

Fig. 5.6, Solution plots for an orifice area ratio of 0.52. ................... ...................... 81

Fig. 5.7, Solution plots for an orifice area ratio of 0.64. ............................................. 82

Fig. 5.8, Solution plots for an orifice area ratio of 0.74. ............................................... 83

Fig. 5.9, Solution plots for the no orifice case. ................................................... 84

Fig. 5.10, Plots of effect of varied spar acceleration added mass term (,s) for an
orifice area ratio of 0.64 ....................... ..... ..................................... 88

Fig. 5.11, Phase plot for linear varying spar acceleration coefficient for an orifice area
ratio of 0.64 .................................................................... .................... 88

Fig. 5.12, Results of solutions calculated with E = 0 and an orifice area ratio of 0.64..... 89

Fig. 5.13, Linear damping only solution results for an orifice area ratio of 0.52. .............90

Fig. 6.1, M odel response with no orifice. ....... ........... ............................................97

Fig. 6.2, Comparison of the model spar and centerwell phase with no orifice ............. 97

Fig. 6.3, Model response with orifice area ratio 0.6......................................................98

Fig. 6.4, Comparison of the model spar and centerwell phase with an orifice
area ratio 0.6.................. ........ ..................................... 98

Fig. 6.5, Chevron Spar response with no orifice based on scaled coefficients
discussed in Chapter 5.................... ......................................................... 99

Fig. 6.6, Comparison of the Chevron Spar and centerwell phase with no orifice ............ 99

Fig. 6.7, Chevron Spar response with an orifice area ratio 0.6 based on scaled
coefficients discussed in Chapter 5. .......................................................... 100

Fig. 6.8, Comparison of the Chevron Spar and centerwell phase with an orifice
area ratio 0.6............................................................ ...... .................... 100

Fig. 6.9, Neptune Spar response with no orifice based on scaled coefficients
discussed in Chapter 5.................................................. 101

Fig. 6.10, Comparison of the Neptune Spar and centerwell phase with no orifice......... 101

Fig. 6.11, Neptune Spar response with an orifice area ratio 0.6 based on
scaled coefficients discussed in Chapter 5........................................... 102

Fig. 6.12, Comparison of the Neptune Spar and centerwell phase with an
orifice area ratio 0.6 ................................................................................ 102


added mass

centerwell cross-sectional area

spar cross-sectional area measured to the outer diameter

orifice area

orifice area ratio AR/Ai

centerwell added mass associated with waves and spar acceleration

spar, wave and effective damping per unit mass

viscous damping coefficient

damping coefficients associated with wave and spar velocities

effective damping coefficient per unit mass

linear or potential damping coefficient per unit mass

added mass coefficients

non-linear or viscous damping coefficient per unit mass

general diameter

centerwell diameter

spar outer diameter

model and prototype outer diameter, respectively






aw, as

Bs, Bw, Be


bw, bs



cm, Ci2, cms





DoM, Dop




h, h, h

h,, ha,,

ho, io,

hi, hi,







n~w, ms


s, s, "S





orifice diameter

orifice diameter ratio DR/DI

wave force acting on a buoy or the centerwell

height of water column within the centerwell measured from the
base of the centerwell and associated time derivatives

position of the centerwell water level measured from still water
level and associated time derivatives
first order solution to the centerwell governing equation and
associated time derivatives
second order solution to the centerwell governing equation and
associated time derivatives

frequency transfer function

amplitude of centerwell water surface motion

hydrostatic pressure coefficient per unit mass

wave number or pipe resistance coefficient

pipe entrance and exit resistance coefficients

general length scale (usually referring to spar draft)

total mass of centerwell and spar

response amplitude operator

spar displacement measured from still water level and associated
time derivatives

amplitude of spar motion

spar draft

model and prototype draft, respectively

damping ratio

6 logarithmic decrement

(4 phase of centerwell water surface motion

rio wave amplitude

77, 77, y water surface displacement due to waves measured from still water
level and associated time derivatives

A(() dynamic magnification factor

ps, P1, [2 added mass coefficient ratios

9 phase of spar motion

On natural frequency (radians/second)

0o reference natural frequency used for comparison purposes

Abstract of Thesis Presented to the Graduate School
of the University of Florida in Partial Fulfillment of the
Requirements for the Degree of Master of Science



William Miller, Jr.

December 1997

Chairman: D. Max Sheppard
Major Department: Coastal and Oceanographic Engineering

This thesis examines the heave response of a spar type platform consisting of a

deep draft buoyant annular structure with a centerwell or moonpool which is open to the

sea at the bottom. The governing equation for the heave response of this free flooding

centerwell is derived using conservation of momentum and its coefficients are calibrated

with experimental data. An analytical solution to this governing equation is obtained and

compared to a numerical solution, giving satisfactory results. Sensitivity tests are

conducted with both the numerical and analytical solutions to explore the effects of the

different components of the centerwell heave motion.

Finally, the governing equation of the centerwell is coupled with that of the spar

platform and numerical simulations are conducted to examine the effects of the centerwell

on the heave motion of the spar platform. The results identify potential design problems

due to the interaction.


In their quest to find and develop new sources of hydrocarbons to feed the world's

constantly growing hunger for fuel, the petroleum industry is pushing offshore into ever

deeper water. The current depth limitations of the various field production options are:

piled steel platforms, 0 to 500 meters, guyed towers, 200 to 600 meters, tension-leg

platforms, 150 to 900 meters, and semisubmersible platforms, 100 to 1500 meters (Clauss

et al., 1992, p. 130). Figure 1.1 shows various production platforms currently in use.

Fixed platforms

SFoating compliant
i platforms

platforms Pile supported
Gravity foundation) jackets Guyed Compliant Articulated
towers piled towers towers Tension leg Catenary anchored
platforms floaters
(tankers, semisubmersibles)

Fig. 1.1, Drilling and Production Platforms (Clauss et. al., 1992)

"A series of new alternative platform concepts are being investigated by the

offshore industry for deepwater development and production prospects. One concept

which is receiving a considerable amount of attention is the Spar Platform" (Niedzwecki et

al., 1995, p. 91). According to Halkyard et al. (1991, p. 41), the Spar Platform, also

known as a Tension Buoyant Tower (TBT), has the potential to extend depth capabilities

to as much as 3000 meters.

The Spar Platform is a large, slender, deep draft, cylindrical floating structure. It

consists of a large, buoyant vertical cylindrical hull topped by a deck structure. The hull

may be used for storage purposes and usually has a centerwell or moonpool. Drilling and

production risers are deployed through this centerwell which also provides them

protection. The deck, which consists of multiple levels, is located well above the waterline

and provides the necessary space for production and drilling operations, pipe storage,

drilling related equipment storage, process equipment, power plant, water treatment

equipment and other facilities. "Its design is intended to eliminate first and second order

heave forces due to the local seaway environment" (Niedzwecki, et al. 1995, p. 41).

Glanville et al. (1991) described a Spar Platform concept designed for Chevron by

Deep Oil Technology and Rauma Repola. The design has a hull diameter on the order of

42.5 meters (140 feet) and a draft of 198 meters (650 feet). The transverse centerwell

area was divided over the upper 50 meters of the hull and had a combined area of 32

percent of the waterplane area of the hull giving it an effective centerwell diameter of 24

meters. The deck structure was located 20 meters above the waterline. The design

operating depth for this platform is 820 meters in the Gulf of Mexico. Figure 1.2 is an

artist's impression of a Spar platform in the Gulf of Mexico. Figures 1.3 and 1.4 are

drawings of a proposed Chevron Spar.

Recently, Oryx installed the world's first production spar (the Neptune Spar) in the

Gulf of Mexico approximately 135 miles southeast of New Orleans. Deep Oil Technology

and Rauma Repola of Finland designed and built the unit. This platform has a 198 meter

draft and an outer diameter of 22 meters and a square 10 by 10 meters centerwell. It is

moored at a location with a depth of approximately 700 meters (http://www.offshore-

technology.com/projects/neptune/ index.html).

"The concept of a deepwater Spar Platform has its origins in the design of

oceanographic spar buoys" (Niedzwecki et. al., 1995, p. 41). The FLIP (floating

instrument platform) Ship was built in 1962 as a stable platform for oceanographic

measurements. Rudnick (1967) documented its favorable motion properties. Newman

(1977, p. 317 321) stated that "FLIP will be stable with practically no heave motion

since its resonance frequency is beyond any predicted wave excitation." In the 1970s Shell

constructed the Brent Spar for oil storage and offloading in the North Sea. Other uses of

the design were discussed by van Santen and de Werk (1976), including a helicopter

landing base, an offshore loading terminal and an oil storage terminal. However,

additional uses of the Spar design have not developed to date.

In the late 1980s and early 1990s, the Spar design was revived and proposed as a

low cost production facility for remote sites. The features which make the Spar attractive

are described by Glanville et. al. (1991) and include: simplicity of design, favorable motion

characteristics, insensitivity to water depth (i.e. the unit may operate in a wide range of

water depths with little variation in the design), ability to support a high deck load, the

centerwell protects the risers from wave loads, the risers can be supported in tension by

means of flotation, low cost of shipyard construction, oil storage availability in tanks

located in the hull, and the structure may be relocated.

Though exhaustive studies have been performed on the basic spar design, the

effect of the centerwell on the dynamics of such a deep draft facility as the Spar Platform

have not been specifically addressed in the open literature.. Informal model tests have

demonstrated a tendency for the water in this centerwell to move synchronously with the

Spar platform at certain frequencies. A small lip located at the bottom of the centerwell

seems to enhance this tendency, even though the flow restriction due to the orifice created

was negligible.

The dynamics of this "synchronous centerwell motion phenomena" are the primary

interest of this study. This paper will attempt to (1) experimentally verify and document

the synchronous centerwell response and determine what effect a bottom orifice has on the

phenomena; (2) examine experimentally and through numerical simulations the dynamics

of the heave motion of the fluid in the centerwell; (3) numerically simulate the combined

Spar-centerwell system in waves and determine potential design problems and/or

advantages due to the presence of the centerwell.

Fig. 1.2, Artist's impression of a Spar platform in the Gulf of Mexico (Randall, 1997, p.

Fig. 1.3, Chevron Spar Platform Mooring Arrangement (Glanville, 1991)

----~.~~- -I~-

Fig. 1.4, Chevron Spar Platform Elevation View (Glanville, 1991)


Spar Buoy/Platform Motion

The heave motion of a spar buoy in waves has been investigated extensively.

Nearly every fluid mechanics and engineering mathematics text uses the system to

illustrate basic principles of mechanical oscillations, buoyancy, damping, resonance and the

use of ordinary differential equations. For the most part, these texts treat the floating

buoy as a mass-spring system and the result is a second order ordinary differential

equation or a system of coupled second order equations, depending on the number of

degrees of freedom. Non-linear terms for drag are introduced in several texts and papers

specifically concerned with buoy design and dynamics.

The basic equation of motion for a floating structure is found by "equating the

external forces on the structure (assumed to be prevented from moving) to the reactive

forces acting on the structure (which is assumed to be moving in a calm fluid)" (Natvig

and Pendered, 1980, p. 99). For a spar type buoy in heave the equation reduces to the

second order ordinary differential equation below. This is the form resulting from the

direct application of Newton's Second Law in an inviscid fluid.

m = a( S)+b(i )+ c(ek r'- s). Eq. 2.1

Referring to figure 2.1, s is the buoy displacement, 77 is the water surface

displacement due to the wave. In equation 2.1, a is the added mass or hydrodynamic

mass, b is the potential damping coefficient and c is the hydrostatic restoring force

coefficient, also termed the spring constant when associating the system with a mechanical




Fig. 2.1, Spar Buoy Arrangement

Equation 2.1 assumes that the wave is long with respect to the buoy and therefore

the buoy is "hydrodynamically transparent." In other words, "there is little change to the

incident wave when it passes the structure" (Clauss et. al., 1992, p. 222). The assumption

is also known as the "long-wave approximation" (Newman, 1977).

Clauss et al. (1992, p. 220) divides the forces experienced by the buoy into "(1)

Froude-Krylov force: pressure effects due to undisturbed incident waves, (2)

hydrodynamic added mass and potential damping force: pressure effects due to relative

acceleration and velocity between water particles and structural components in an ideal

fluid, and (3) viscous drag force: pressure effects due to relative velocity between water

particles and structural components."

Newman describes the Froude-Krylov hypothesis as "the assumption that the

pressure field is not affected by the presence of the body and can be determined from the

incident wave potential by itself. The diffraction potential is neglected completely..."

(1977, p.305). Newman goes on to state that, given the wavelength is large compared to

the beam, the Froude-Krylov exciting force and moment are the leading-order

contributions in the vertical plane.

Assuming a linearized form of the well known Bernoulli equation, Clauss et al.

(1992, p. 265) describes the Froude-Krylov force in equation 2.2, where 0o is the incident

wave velocity potential. Newman (1977, p. 299-309) ), Hooft (1982, pp.144-153) and

Sarpkaya and Isaacson (1981, pp. 441-443) develop similar equations and results.

FFK = -P I pndS =p S Eq. 2.2
s, s, at

For a spar buoy in heave due to harmonic wave motion, the Froude-Krylov (FFK)

force is reduced to the wave portion of the hydrostatic restoring force which in deep water

is FF = pgAek"l= cek q (Newman, 1977, p. 357-359; Hooft, 1982, p. 144-151; Clauss

et al., 1992, p. 265-268).

Newman (1977, p. 289-296), Sarpkaya and Isaacson (1981, pp. 438-440), Hooft

(1982, pp.112-114) and Clauss et al. (1992, pp. 288-291) demonstrate that near a free

surface, the hydrodynamic added mass and potential damping force can be derived from

the dynamic part of the unsteady Bernoulli equation. Applied to the velocity potential of

the waves generated by the oscillating body, this results in equation 2.3, where Se is the

equilibrium body surface.

p J -dS = -(a + b) Eq. 2.3
s, Ot

Substitution of equations 2.2 and 2.3 into the equation for Newton's second law of

motion results in an equation of the form of equation 2.1. This may be rearranged to give

equation 2.4 in the common mass-spring system notation for a shallow drafted buoy.

(m+a) +bs+cs= ai+bi+cyq= f(t) Eq. 2.4

This is the simplest form of the equation of motion for a buoy in heave derived in

basic design and dynamics texts (Korvin-Kroukovsky,1955; Lewis, 1967; Berteaux, 1976;

Newman, 1977; Hooft, 1982; Patel, 1989; Clauss et al., 1992; etc.). Many of the authors

perform the derivation with the coefficients expressed as matrices and the variables as

vectors to develop two and three dimensional coupled systems of equations and apply

equation 2.4 to up to six degrees of freedom. However, most authors studying multiple

degrees of freedom limit themselves to heave and pitch. The equation of concern in this

paper is limited to a single degree of motion since "the heaving motion.., mode of

oscillation is of prime interest with spar type structures" (Kokkinowrachos and Wilckens,

1974, p. 99). As will be noted later in this chapter, Aalbers (1984) in studying the motion

of water in a moonpool also choose to limit his investigation to heave.

For deep drafted buoys and structures, a wave "attenuation factor" should be

applied to the wave force terms in equations 2.1 and 2.4 to account for the variation of

water particle motion with depth (Berteaux, 1977, p. 54, and Lewis, 1967, p. 629). This

attenuation factor is derived directly from linear wave theory and in deep water the factor

reduces to ekz, where k is the wave number (for deep water k = w2 /g, with w the wave

radian frequency) and z the depth (negative downward) (Dean and Dalrymple, 1991). For

very deep draft structures where the "draft (So) is large compared with z, the factor

becomes approximately ek = ekSo." (Lewis, 1967, p. 629). If it is assumed that most of

the wave forces act on the bottom of the buoy, the application of the attenuation factor

results in equation 2.5 below.

(m+a)s+bs+cs= e-kSo (ai+bil+cek7) = f(t) Eq. 2.5

As previously noted, the "linear damping" term represents the "potential damping"

due to the body moving through an inviscid fluid. Glanville et al. (1991, p. 61) refer to

this damping term as "linear damping... due primarily to radiation," alluding to a

dependence on wave generation by the buoy or platform.

The potential damping may be obtained theoretically by integrating the pressure

distribution over the body based on the potential of the undisturbed wave (Vugts, 1968a,

p. 16). This procedure is known as "the linear diffraction problem" and should be used

when "the structure spans more than about a fifth of the incident wave length" (Sarpkaya

and Isaacson, 1981, p. 382).

However, Newman (1962 and 1977, p. 303-307) favors a method due to Haskind

(1957) known as the "Haskind Relations, which express the damping coefficients in terms

of the exciting forces." Therefore, an "actual solution to the diffraction problem can be

avoided" (Vugts, 1968, p. 16). This expression is endorsed by Garrett (1971, p. 129),

Vugts (1968, p. 16; 1971, p. 21), Kokkinowrachos and Wilckens (1974, p. 107-108),

Sarpkaya and Isaacson (1981, p. 440-445) and Hooft (1982, p. 121-124) for motion in the

presence of a free surface. Newman derives the relation and notes that it is not dependent

on the long-wave approximation. The resulting equation is identical among the authors

and is

k 2 X,(0)2
bii gC 21X() 2 dO. Eq. 2.6
87rpgC, o 7,

In equation 2.6, k is the wave number, Cg is the wave group velocity, 7ro is the

wave amplitude and X, is the complex amplitude of the exciting force. Assuming a single

frequency linear wave of the form 77= ro cos(cot) and returning to the long-wave

approximation, X, can be derived from equation 2.5. The result in complex form is

X, = e o(-aco2 +c-ibco) Eq. 2.7

Hooft (1982, p. 121) indicates that the equation should be evaluated for each wave

frequency in the incident wave spectrum. Integrating the equation for an axial symmetric

structure with a vertical axis, and assuming deep water linear waves, leads to the following

equation for the linear damping coefficient in heave:

S03 2
b .X Eq. 2.8
2pg3 I

However, the problem of the damping term does not end there. The linear system

of equation 2.5 has neglected the viscous effects. For large manned platforms this may be

sufficient. "Due to the large diameter of such structures the potential flow linear solution

is normally of primary importance while in this case viscous effects are negligible"

(Kokkinowrachos and Wilckens, 1974, p. 99).

However, many authors include these viscous effects and when doing so "it is

advisable to stick to the simple structure of the Morison-formula" (Kokkinowrachos and

Wilckens, 1974, p. 105). The Morison equation assumes that the total wave force is "the

sum of the inertial force and the viscous force" (Newman, 1977, p. 42) where the viscous

force is proportional to the square of the velocity. Including such a "viscous damping"

term results in equation 2.9.

(m + a)> + (bea )s+ (bnoninea ) + cs = f(t). Eq. 2.9

This is the equation given by Hooft (1970, p.35), Ochi and Malakar (1984, p. 5)

and Glanville et al. (1991, p.61). Snyder and Darms (1967, p. 151), Gudmestad and

Connor (1983, p. 184), Patel (1989, p. 316) and Clauss et al. (1992, p. 278) express the

viscous damping in terms of the relative velocities between the spar and the wave particles

as was done for the potential damping in equation 2.1 resulting in equation 2.10.

(m + a) + (bnea)s + (boninear) ) s- + cs = ai + (binea )il + (bnoni,, )rI 77 + cr .

Eq. 2.10

To develop a linear approximation to equation 2.10, Gudmestad and Connor

(1983, pp. 185-186) suggest that for >> i, the (bnon,,)il-(il-s) term may be

replaced by (bnon ,,)ear) j (b noninear )S

Patel (1989, p. 316) adds (bnonie,)as to both sides of the equation to obtain a

(bnonlinear) ) s (i- s) + bnoninear term in the wave force expression. He then states that

the "relatively large magnitude of the inertia term in the wave force and the decay with

depth of wave particle velocities... makes the" (bnonlinar,)]l- s\(l- s)+bnoniInear s "term

small for members with significant submergence below still water level." Thus, for


(bnoni,,ear )li7 s-(i s) +bnonlnea, Is -(b noninear ) + bnonunear, I = 0

Then, for small il, the (bnoninr) I-l( si- ) +bnolinea, ss term may be replaced by

(bnonlinear)liil and equation 2.10 can become equation 2.9 which is somewhat more

manageable. Performing this expansion results in equation 2.11.

(m + a) + (b,,,) + (b ,onln )\s + cs = [ai + (b,,I)il+(bnonnea,) l + c ]e.
Eq. 2.11

Snyder (1967, p. 152) and Hooft (1970, p. 35) describe the classical determination

of the drag force on a structure where the non-linear damping coefficient can be defined in

terms of the "area of projection of the body in the direction of velocity" (Ap) and a drag

coefficient (CD) as given in equation 2.12.

bnon-inear = PCDAp. Eq. 2.12

Snyder and Hooft refer to the experimental work of Hoerner (1965) for values of

the drag coefficient. Hoerner explains that "the drag on a smooth cylinder in supercritical

(flow) condition (Reynolds number greater than 106) is approximately constant" (1965, p.

3-11 to 3-13). Hoerner plots experimental results for "cylindrical bodies in axial flow,

with blunt shape" and gives a value of total CD for viscous pressure drag at approximately

0.85 for length (draft) to diameter ratios greater than 2. "The drag of these shapes

essentially consists of that of the forebody and base drag." It should be noted that these

drag coefficient values are not those for a body in oscillatory motion.

Snyder (1967, p. 153), corrects for oscillatory motion by dividing the non-linear

drag coefficient itself into forebody and base drag components, forebody drag defined as

the form or pressure drag due to the body's shape and the base drag essentially as the skin

friction component. He states that the drag coefficient (Co) "for surface piercing buoys...

changes with the direction of relative flow. When the buoy is moving downward into the

water the forebody drag coefficient must be used. When the buoy is moving upward with

respect to the water, the base drag coefficient must be used." From Hoerner (1965), the

base drag coefficient alone is approximately 0.2 for length to diameter ratios greater than

4 and up to 12.

Many designers and researchers linearize equation 2.10 or 2.11 by developing an

"equivalent linearized damping coefficient" in accordance with a method developed by

Kryloffand Bogoliuboff(Berteaux, 1976, App. I; Dao and Penzien, 1982; Gudmestad and

Connor, 1983, p. 189; Tung, 1986; Patel, 1989, pp. 290-291 ; Ochi, 1990, pp. 389-409;

Clauss et al., 1992, pp. 278-279). The principle of the Kryloff-Bogoliuboff method is to

"consider an equivalent balance of energy during one cycle such that work per cycle of the

two systems (non-linear and equivalent linear system) is the same" (Ochi, 1990, p. 389).

Thus, the procedure is to express the energy dissipation for linear and non-linear damping

over a cycle in terms of velocity, assume harmonic motion and equate the results and solve

for the linear damping coefficient in terms of the non-linear coefficient. By integrating the

\^s term over a cycle the equation may be linearized using the approximation

sin(cot) sin(ct) = (8/3r) sin(at).

Berteaux, Dao and Penzien, Gudmestad and Connor, Ochi and Tung derive their

equivalent term as a frequency dependent term. Patel develops his expression in terms of

the maximum energy dissipated and thus it is only dependent on the natural frequency of

the system. Clauss et al. express the equation with an implied frequency dependence with

the result being a function of the amplitude of the relative vertical velocity. Additionally,

Ochi and Gudmestad and Connor explore several other linearization methods with special

emphasis on the application of the resulting response transfer functions to random wave

spectra. Several of the expressions for linearizing the damping term in accordance with

the Kryloff-Bogoliuboff method are shown below.

8 8
Berteaux: btinear = (bnonlinear)a = -(bnn-linea,)R Eq. 2.13

Patel: bia= (bnon-r) = (non-inear)RoS Eq. 2.14
37c 31r

Clauss: blinear = (bno,,inar)- Eq. 2.15

Ochi: binear = (bnon-inear)O- Eq. 2.16

In equations 2.13 through 2.15, s\ =Rco is the amplitude of the structure's

velocity for frequency c, and m. = Raco is the amplitude of the maximum response

velocity of the structure, which occurs at the natural frequency, i7 is "the vertical orbital

velocity" of the water particle "at the hull center and s is the vertical velocity of the

heaving semi-submersible" and I a "is the associated amplitude of relative vertical

velocity" (Clauss et al., 1992, p.278). In equation 2.16, -, is the standard deviation of

the response velocity (Ochi, 1990, p. 391; Gudmestad and Connor, 1983, p. 189).

Clauss et al. expand this amplitude term with respect to the complex forms of the

harmonic wave and spar motion resulting in the approximation of equation 2.17 (zhc is the

hull center where the water particle velocity is taken).

binear = bnon-inear 8- = T (bnonear )o e kz -- Eq. 2.17

The dependence on response amplitude require that the linearized equation of

motion must be solved iteratively to determine a response amplitude value that satisfies the

corresponding equation. Additionally, equation 2.17 requires that the phase also be


The hydrodynamic mass or added mass of the system is another frequency

dependent term. Sarpkaya and Isaacson (1981, p. 440) note that "it should be emphasized

... that both the added-mass and the damping coefficients are frequency dependent."

Kokkinowrachos and Wilckens (1974, p. 108) use a similar derivation to the Haskind

Relations and assert a dependency of the added mass on the excitation force. Hooft

(1970, p. 24-26) states "in general these coefficients ... depend on frequency (and) when

the body has a three dimensional form ... the added mass is not only influenced by the

projected area in the direction of oscillation but also by the length of the body in the

direction of oscillation."


Sabuncu and Calisal (1981) performed a theoretical analysis of the added mass

and damping coefficients for the motion of a vertical cylinder which included varied draft

to diameter ratios. Figure 2.2 shows the deep water results for the added mass. In this

figure T is the draft of the cylinder and A is its radius. The added mass coefficient

corresponds to that in equation 2.4. Sabuncu and Calisal (1981, p. 60) state that "the

results are compared to some available experimental numerical results. The agreement

with numerical results are observed to be satisfactory." The figure shows that as the draft

to radius ratio increases, the added mass coefficient becomes smaller and less dependent

on frequency.

Hydrodynamic coefficients for vertical circular cylinders at finite depth

Vertical cylinder heave deep water
A T/A-0.1
T T/A- 0.3
9.6 T
._ _a T/A-0.5
.--+A T/A-3
t 7.2-


2.4 >

0 1.2 2.4 3.6 4.8 6
Omega Omego* A/G

Fig. 2.2, Added mass coefficients for heave motion of a vertical cylinder, G = gravity, A =
Diameter, T = draft (Sabuncu and Calisal, 1981, p. 41)

Hooft (1982, p. 150), Patel (1989, p. 286) and Clauss et al. (1992, p. 279-283)

recommend values of constant added mass coefficients in their discussions of buoy and

platform design. Clauss et al. incorporates a form coefficient to account for the draft to

radius ratio. Figure 2.3 shows the values recommended by Clauss et al. (ms3 is the added

mass for heave motion). Patel recommends the added mass of a half hemisphere of the

same radius as the cylinder and Hooft recommends using the added mass of a circular disc

with this radius .

added masses mkj
(direction of force k i direction of motion j)
Ia 7 hulls with rectangular cross-section
m,33R Co(aUR) nb2 (length I)
1.8 caissons columns

S2 s cs m3c s=C(q jPR, mC3 c=Cjac)PR
1 26
Sfor comparison: horizontal circular cylinder

et al. 1992, p. 281)(C1)

o 1.2____


0 1 2 3 4 5 6 7 8 9 -10
UR= b 'C' =ARCS .actRC
a 2hcs C4h I
Fig. 2.3, Added masses of vertically oscillating components of offshore structures (Clauss
et al., 1992, p. 281)

The effect of added mass may be used to minimize the heave of a buoy attaching a

plate to the bottom of the buoy (known as a "damping plate"). The idea is to "minimize

the heave by adding to the virtual mass of the buoy and also to provide some damping

through the drag of the plate" (Cavaleri and Mollo-Christensen, 1981, p. 24). Cavaleri

and Mollo-Christensen found through numerical simulations that for higher frequencies the

damping plate was effective, but for lower frequencies it may actually lead to increased


The hydrostatic portion of the pressure term is called the hydrostatic stiffness

force, the buoyancy force, the restoration force or even the spring force when relating the

system to a mechanical oscillator. This force results from "the total change in buoyancy

from the initial calm-water condition as a result of both the change in water level" due to

the wave "and vertical displacement" of the buoy (Lewis, 1967, p. 628). The hydrostatic

force coefficient or spring constant is proportional to the waterplane area (Aw) of the

buoy: c = pgA, (where p is the fluid density and g is gravitational acceleration).

Figure 2.4 illustrates a composite spar platform of the type used for storage and

loading in the Brent Field in the North Sea. By proportioning the structure so "the length

of the (surface piercing) column is such that the superstructure is high enough out of the

water and that the main body is out of the wave zone" the design minimizes the wave

forces experienced by the platform (van Santen and de Werk, 1976, p. 1106).

Fig. 2.4, Composite Spar Platform

Structures such as the composite spar are classified as semi-submersibles. Figure

2.5 shows examples of semi-submersible platforms classified as "multi-hull column

stabilized structures" consisting of "an array of vertical columns of minimal waterplane

area, which are connected to caissons or longitudinal hulls with interconnecting truss-like

structural members below the water surface" (Clauss et al., 1992, p. 76-77). The

centerwell designed into the Spar Platform described in Chapter 1 would appear to have

the same advantage of minimizing the waterplane area of the platform while providing the

deep draft required to minimize wave forces.

PACESETTER (1973) AKER H-3(1974) BINGO 3000 (1982)

GVA 4000 (1983) TRENDSETTER (1987) BALDER (1978)
Fig. 2.5, Examples of Semi-Submersible Platforms (Clauss, 1992)

Effects of a Moonpool or Centerwell

Garrett (1970) theoretically investigated the excitation of waves inside a partially

immersed open circular cylinder. The cylinder examined was thin walled and floating in

water of finite depth. Garrett developed his solution using the linear diffraction method

PENTAGONE (1969) SEDCO 700 (1973)

SEDCO 135 (1965)

and solving the continuity equation inside and directly below the cylinder. He concluded

that (1) the phase of the response was independent of depth, (2) resonances occurred at

wave numbers close to those of free oscillations in a cylinder extending to the bottom and

(3) away from resonance, the response amplitude was less then 10% of the incoming wave

amplitude if the draft of the cylinder is greater than /4of the wave length of the incoming

wave (p. 433). Garrett also noted that the sharpness of a resonance depended on the

value of the ratio of the cylinder draft to incoming wave length, with the cylinder of

greater draft producing the sharper resonance.

Mavrakos (1985) continued Garrett's work by theoretically examining a

stationary, floating, bottomless cylinder with a finite wall thickness in a finite depth of

water. By dividing the fluid into three regions (outside the cylinder, below the cylinder

walls and inside and below the centerwell) and then solving the continuity equations in

these areas, Mavrakos conducted numerical experiments to predict the motion of the fluid

inside the cylinder. He concluded that the ratio of the outer cylinder radius to the

centerwell radius had "practically no influence" on the amplitude of the response of the

fluid in the centerwell. He also noted that increasing the outer radius to inner radius ratio

for a constant depth to inner radius ratio tends to narrow the width of the resonance


Mavrakos (1988) went on to examine the forced motion of a floating, bottomless

cylinder with finite wall thickness in finite water depth. Again, starting from the continuity

equations and using a linearized Bernoulli equation to determine pressure, he derived

equations for the hydrodynamic coefficients in finite water depth. As these equations are

taken into deep water, the damping coefficient vanishes while the added mass blows up.

However, Mavrakos did note that resonance situations arise in the centerwell at

frequencies corresponding to the wave lengths of free fluid motions in a cylindrical

container with a bottom.

All of the above studies were theoretical and all used linearized approximations.

Knott and Flower (1980) performed experiments in which energy losses were measured

for oscillatory flow through a pipe exit with various shapes and curvatures and used

photographic visualization to examine eddy formation near the exit. Dividing the losses

into linear damping and non-linear "eddy damping," they found that the linear losses were

accounted for by the theoretical viscous friction of the pipe walls and the eddy losses were

dependent on the radius of curvature of the exit, the radius of the tube and the oscillation

amplitude and frequency. From the flow visualizations it was found that "eddies formed

even at low amplitudes (of oscillation), creating a local turbulent field below the mouth,"

however "eddy losses only begin to act once a certain amplitude of motion has been

exceeded, and that critical amplitude is strongly dependent of the radius of the lip." (p.

163). "Photographic evidence... suggests that in some cases where high levels of

turbulence occur on the outflow the induction of eddies back into the mouth appear to

cause a blockage with necking of the flow and high shear stress gradients near the walls"

(p. 160). Knott and Flower also concluded that "geometrical similarity was the most

important consideration... since this ensured similarity of the dominant parameters" and

that "larger scale operation involving higher Reynolds numbers would, if anything, suffer

less eddy losses owing to the delayed separation of boundary layers" (p. 164).

Aalbers (1984) analyzed the water motions measured in the moonpool of a model

ship. Based on the theory of Van Oortmerssen (1979), which describes a system of two

floating structures in waves, Aalbers assumed that "the moonpool water column oscillates

in such a way that it may be replaced by a frictionless piston inside the vessel" (Aalbers,

1984, p. 578) and developed the equations of motion accordingly. His equations are

shown below.

For the moonpool:
(pA(T + h) + ah } + bh + b2hh + pgAh
+ (d, + pA(T +h) +ah, + (eh +bh) + pgAs Eq. 2.18
+higher order terms = F .

For the vessel:
fM + a, }I + b,s + cs + dh (N + h+ e + Eq. 2
{M+aS}l+bs~+c~s+d,(j+h(+h)+esh() Eq. 2.19
+higher order terms = F .

In equations 2.18 and 2.19, s is the vessels vertical displacement (heave), h is the

relative displacement of the surface of the moonpool (absolute motion is the sum of the

ship's heave and the relative motion), T is the vessel's draft, ah and a, are the added mass

of the moonpool and vessel (from potential theory), bh and bs are the damping coefficients

of the moonpool and vessel (also from potential theory) and b2 is the quadratic or vortex

damping coefficient.

Van Oortmerssen investigated the interference effect of neighboring vertical

cylinders and developed the hydrodynamicc interaction coefficients" (ehs, eh, dhs, dsh) to

describe the forces experienced by one body due to the motions of the other body (1979,

p. 342). The coefficients are symmetric, therefore ehs=esh and dhs= dsh.

Finally, Fwh is the exciting wave force. The higher order terms are due to

Bernoulli's expression for fluid pressure and are neglected by Aalbers (1984, p. 558).

Aalbers also noted that the moonpool can be excited at frequencies equal to one half the

natural frequency due to the non-linearities.

Aalbers experimented with three moonpools with diameters of 0.1, 0.2 and 0.3

meters. The model's draft ranged from 0.15 to 0.3 meters with a maximum draft to

moonpool diameter ratio of 2. Tests were conducted in water depths equal to the model

draft plus 0.6 meters. The results showed a "dependency of the added mass and both

damping coefficients on the size of the moonpool, especially on the diameter." It was

determined that the added mass was proportional to the diameter cubed (ah c dia3), the

linear damping was proportional to the diameter to the five halves power (bh ocdia2),

which agrees with Froude scaling, and the quadratic damping was proportional to the

diameter squared (b2 oc dia2) (Aalbers, 1984, p. 561).

Large Spar Platforms

The FLIP (Floating Instrument Platform) manned spar buoy was a large spar

platform used for oceanographic research in the 1960s. FLIP had a draft of

approximately 91.5 meters and a tapered hull with a diameter ranging from 6 meters at the

keel to 3.8 meters at the surface. Rudnick collected and analyzed motion information

onboard FLIP at sea and found that the maximum response amplitude to wave amplitude

ratio was 0.177 at a frequency of 0.059 Hz (Rudnick, 1967, p. 264).

Glanville et al. did a preliminary motion analysis study on the Chevron Spar design.

Their analysis was conducted using the JONSWAP spectrum and found a low frequency

response with a signal to wave response of 0.24 at a frequency of 0.043 Hz (Glanville et

al., 1991, p. 67). They noted that "in general, the computed response agreed with the

results of the model experiments, including the low frequency effect" (p. 62).

Finally, Niedzwecki et al. (1995) conducted experiments on scale models of the

Chevron Spar Platform described in Chapter 1 to investigate the surge motion of a

tethered structure. One of their results was that "Morison's equation can produce

satisfactory results for inertia dominated structures with diameter to wave length ratios

less than 0.2 if it is accompanied by the appropriate modifications (for non-linearities)"

(Niedzwecki et al., 1995, p. 105).


Momentum Equation for Heave

Figure 3.1 shows the spar and centerwell configuration for which the equations of

motion will be derived. It will be assumed that the centerwell air water surface has a

uniform, level elevation. Two centerwell levels are of concern: the relative centerwell level

or depth (h) measured from the bottom of the spar (and always a positive value) and the

absolute centerwell level (ha) which is the displacement of the centerwell water surface

measured from a fixed horizontal datum (the still water plane).

The centerwell is bounded by the spar inner diameter, DI, and has a cross-sectional

area A = nDI The spar outer diameter, Do, also has a corresponding cross-sectional

area, Ao = -nD Finally, an orifice plate may be placed at the bottom of the centerwell

with an orifice diameter, DR, and area, AR = 4nD2 A useful non-dimensional parameter

is the ratio of the orifice area to the centerwell area referred to here as the "area ratio,"

AR = AR/A,.

The control volume used in the derivation of the governing equation is shown in

Figure 3.2. The control volume (CV) is the fluid in the spar (i.e. the centerwell). It is

attached to the spar at the bottom and its upper surface moves with the centerwell surface.

Thus the control volume varies with time and has a volume V = hAI.



spar still ......* h
waterline ...... ........... ......
So = still water draft
s = SPAR displacement
centerwefrom reference
h = centerwell depth
So s S, from bottom of spar
ha= centerwell surface
displacement from
orifice lip
\l = wave displacement


DRFig. 3.1, Spar and Cente ell Configuration and Definition Terms (not to scale)
Fig. 3.1, Spar and Centerwell Configuration and Definition of Terms. (not to scale)



Fig. 3.2, Control Volume and Control Surface for deriving the Centerwell Equation of
Motion. (not to scale)

The integral form of the conservation of momentum equation for the control

volume is:

F= JpdV + f p(Vr fii)dA Eq. 3.1
at cv C s

In equation 3.1 p vis the linear momentum in the differential volume, dV, at any

time and pv(v", -n)is the flux of momentum through the differential surface area, dA.

Aalbers modeled the moonpool motion by assuming it was a frictionlesss piston." The

draft to centerwell diameter ratio for the Chevron Spar and thus the experimental model is

approximately 8, while that of the vessel examined by Aalbers was 2. Thus, based on

Aalbers success and the large draft to diameter ratio it is assumed that the fluid motion in

the centerwell is restricted to the direction of the spar axis and that it has a uniform

velocity equal to ha.

With the above assumption and noting from Figure 3.1 that h = SO -s+ha, the

control volume term of equation 3.1 becomes the expression in equation 3.2.

d (jfV = p ( f kadV = f 4 aAh) =Pd [ aA (S s+-ha)]
dt c -v a c vdt d
Eq. 3.2

Taking the derivative and recognizing that both s and ha vary with time while So is

constant results in equation 3.3.

SpdV = pA [ha +ha (a -) = phAha +pAh pAh
Eq. 3.3

Next consider the flux term of equation 3.1. Since the relative velocity of the

upper surface of the centerwell control volume is zero, there is no flux of momentum

through this surface. The only surface through which fluid and momentum may flow is the

open surface at the bottom which is bounded by the orifice. The unit outward normal

vector is downward on this surface. Therefore, if it is assumed that the flow is uniform

and in the direction of the spar axis, vr, = -v', where VR is the average relative velocity

through the orifice.

Applying the continuity equation to this control volume and this control surface

gives the following:

Vr -Vr A AR

This equation expresses the relative velocity over this portion of the control

surface. To find the flux of momentum of the fluid, the absolute velocity must be

determined. Under the same assumptions (i.e. uniform velocity in the direction of the spar

axis) the momentum per unit volume of the fluid in the orifice is given by:

AT = (A, A,
( AR AR AR )

Thus the flux term in the momentum equation becomes equation 3.4,

jpV(V, A = p + ha -hA Eq. 3.4

Integrating over AR (the area of the orifice) yields equation 3.5 below:

fA -)dA =ApA AA A. A -s

= A (p 2 ~+2 ,I 2 Eq. 3.5
AR a AR a AR a AR)

Combining equations 3.3 and 3.5 results in equation 3.6,

zF =pAh a +ha -h pA2i -ha A h +2 A, I2

=PAIhh +PAIhA 23ha + 2 (2 -2ha +2) Eq. 3.6

Finally, the conservation of linear momentum for the centerwell control volume

reduces to equation 3.7.

YF=pAIha +pPA4 l- (ha-)2 =pAhha +pA (1- Al 2. Eq. 3.7

Inviscid Forces on the Fluid in the Control Volume

Next consider the inviscid forces acting on the fluid in the control volume. The

forces can be divided into body forces (FB) and normal surface forces (Fp) as shown in

equation 3.8 below:

ZF=F+F + = pfdV- J pdS, Eq.3.8
cV S

where f is the body force per unit mass, p is the pressure and S is the surface area of the

centerwell control volume.

If the only body force considered is gravity, then equation 3.8 becomes:

F = FB + Fp = -pgAhh + I phdS. Eq. 3.9

Due to symmetry the pressure forces on the vertical sides of the centerwell have a

net value of zero. The pressure forces on the upper and lower surfaces of the centerwell

are considered separately. On the upper surface, the pressure is atmospheric thus the

integral over this surface is:

f pidS = pa,,Az. Eq. 3.10
upper surface

At the bottom surface the pressure has both hydrostatic and dynamic components.

The hydrostatic component is time dependent due to (1) the position of the spar base

varying with time and (2) the time variation of the water surface due to waves. The

variable dynamic pressure is the result of the relative motion between the base of the spar

and the surrounding fluid. As noted in Chapter 2, Sarpkaya and Isaacson state that a

structure must span more than approximately one fifth of the incident wave length before

the system enters the diffraction regime where the long wave approximation does not

apply (Sarpkaya and Isaacson, 1981, p. 382). For a structure such as the Chevron Spar

with a 43 meter beam as described by Glanville et al. (1991), this corresponds to a

minimum deep water wave period of approximately 12 seconds. Energy spectra for deep

ocean waves show that the majority of deep ocean wave periods fall within a range of 2.8

to 21 seconds (Ochi, 1997) and the 100 year storm modal period used in the JONSWAP

spectrum by Glanville et al. (1991, p. 65) for their analysis of the Chevron Spar was 14

seconds. Therefore, though near the limits of applicability, the Froude-Krylov hypothesis

and the long wave approximation should be adequate for the present study.


Fig. 3.3, Spar diagram to scale.

Consider the spar illustrated in Figure 3.3. This figure is a scale diagram of the

spar model used in this thesis which is geometrically similar to the Chevron Spar. The

draft to centerwell diameter ratio for this spar is approximately 8 and the orifice area is

one half of the total centerwell cross-sectional area. This has been arbitrarily set as the

minimum orifice area for this analysis.

Given the relatively small size of the orifice lip in comparison to the centerwell

depth, it was assumed that the orifice lip imparts no significant acceleration to the

centerwell fluid, but rather serves to accentuate the turbulent eddies which Knott and

Flower found to form at the mouth of a pipe under oscillating flow (Knott and Flower,

1980). Therefore, corrections for these flow variations due to the presence of the orifice

lip will be included in the non-linear viscous damping coefficient. The validity of this

approximation can be determined by how well the derived equations predict the

experimental results.

Assuming deep water waves and using the above approximations, the pressure on

the lower surface of the control volume is given by the expression in equation 3.11. The

hydrostatic portion is assumed to be due to the water level outside of the spar.

Power s = ce 2h -g(So s +e" k]+Pam. Eq. 3.11

Integrating equation 3.11 over the centerwell cross-sectional area (Ai) at the

bottom of the control volume and recognizing that the unit normal to this surface is

downward results in equation 3.12.

I pldS = f ( s + Re Pkzj lS
A, A[ a 2 ] Pat
ApidS=J p- h -gSO + -1)dS

=plf dS+lph2A, +pg(So -s+ e")A, -ptmA,. Eq. 3.12
AI 01 2

Using the relation described in equation 2.3, the unsteady term may be reduced to

the in-phase and out-of-phase terms below (note the potential term includes both the

centerwell velocity potential, the spar velocity and the incident wave potential):

p fJ S = -a,, -ie-bh-e -a --bh -S), Eq. 3.13

where a is the added mass and b is the potential damping coefficient. The results are

separated into terms involving the relative velocity and acceleration between the

centerwell fluid and the wave and between the centerwell fluid and the spar, thus leading

to hydrodynamic coefficients for the wave (aw, bw) and the spar (as, bh).

Incorporating equation 3.13 into equation 3.12 and combining the result with

equation 3.10 results in the pressure force given in equation 3.14.

f pfidS = patmA, -a. (a-h e)-bw ha-elek)-a, (fa -s)-b ha-s)
+ ph A, + pg(S, s +re)A -PtmA

= (a, +as,)ha -(hb, +b,)ha +a, S+bs phaA, +pg(S, -s)AI
+ekz(ai + b, + pgA, ) .
Eq. 3.14

Substituting equation 3.14 into equation 3.9 and recalling that h=So-s+ha yields

the final form of the inviscid force term in equation 3.15, where a, = a, +a, and

bT = b +b,.

F F=-pgA,(S, s +h)-aTh r -blha +a,N +b,s +- ph A, +pg(So -s)A,
+ek (aij+b, + pgAi7)

= -(ah +bha +pgAha ph A) +a,s+bh,+e' (awi+bw.+pgA,7).
Eq. 3.15

In equation 3.15, the added mass and potential damping coefficients should be

dependent on frequency. However, as shown in Figure 2.2, the added mass is nearly

constant with frequency for the draft to diameter dimensions of this centerwell (Sabuncu

and Calisal, 1984).

Viscous Damping

Equation 3.15 does not include the damping effects due to skin friction and vortex

generation at the bottom of the centerwell. Aalbers (1984, p. 578) and Knott and Flower

(1980) found that a non-linear viscous damping term was required to account for these

effects. Therefore, a viscous damping force dependent on the relative velocities between

the spar and fluid is introduced as given in equation 3.16.

Fv = -b, ha -s- -s- ). Eq. 3.16

Adding this force term to equation 3.15 results in equation 3.17.

F = aha +(bT +b, ha -s-eki) a +pgA,h, I phA +a,

+ (b, +b ha --e"k' t +ek [ai+ (b +bd ha --ekz 1)i+pgA,1].
Eq. 3.17

Complete Equation for Heave Motion

Combining equations 3.7 and 3.17 and rearranging terms and dividing by the total

mass (pAh+a, = pA,h(l+cm +cm2) = pAh(1+Cmr)) results in equation 3.18, where Cmr

is the total added mass coefficient, Cm is the wave added mass coefficient (a, = cApAih)

and c,2 is the added mass coefficient for the spar (a, = cm2pAzh).

pAihha +pA 1- -h2 [aTa + (b +b, h -s-e h)a +pgA,h, -Ppha A1

+ a,+ (b, +(b,+,h,-a e s+e" [a. + (hb +b ,l, h-S ke")) +pgA, 77]

b +b a -s-e g 1 12 f A
ha + ha + g ha h2a- 1 2=
pAh(1+cmr) h(1+crT) h(l+ CmT) 2 AR

cm2 + (bs,+bh -.-eki)
S+ pAh(l + CT )
l+ Cmr pAih(1+cmr)
S(b. +blka -S-eki
1+cr pA, h(l+c,,r h(l+Cm)

Eq. 3.18

Recall Aalbers equation for the moonpool fluid motion in Chapter 2 (equation

2.18). The correspondence between the terms in Aalbers' equation and the parameters

shown in Figure 3.1 are given in table 3.1.

Converting Aalbers' terms in equation 2.18, and rearranging the terms and dividing

by the total mass of the moonpool results in equation 3.19.

b b +- a + g h
pAh(l+c,) pAh(1+cr,) h(l+cr)
Eq. 3.19
b order terms F
+higher order terms = s- s + Fwh
pAh(l +cer) 1+c r

Table 3.1, Comparison of term in Aalbers' equation (Aalbers, 1984)
and the parameters in Figure 3.1.
Aalbers term Figure 3.1 term
h ha-s
(moonpool surface displacement with
respect to vessel still water line relative
to the vessel)
T+h h
(moonpool depth)
bh b
(potential damping coefficient)
bz by
(viscous damping coefficient)
A Ai
(moonpool cross-sectional area)
dhs, ehs as, bs (possible equivalent terms)
hydrodynamicc interaction coefficients)

Converting Aalbers' terms in equation 2.18, and rearranging the terms and dividing

by the total mass of the moonpool results in equation 3.19.

h +- h +- -b + g7h,
pA,h(l+c) pAh(1+cm) h(l +c)
Eq. 3.20
b c
+ higher order terms = 1 h(1- Cm Fh .
pAh(l+cT) 1+cmr

Aalbers modeled the moonpool as a separate structure, a frictionlesss piston," and

used the method of Van Oortmerssen (1979) to develop hydrodynamicc interaction

coefficients" to account for the interaction between the moonpool and the vessel.

Equation 3.18 does not use Van Oortmerssen's method. Even though the centerwell fluid

is assumed to move in a uniform manner, it is not treated as a separate structure in this

thesis. Rather, the centerwell is modeled as a fluid coupled to the spar through viscous

and normal forces, which both excite the motion and restrict it. Equation 3.18 contains

terms which are similar to (and are here assumed to be equivalent to) the hydrodynamic

interaction terms of Aalbers.

The other major difference between equation 3.18 and equation 3.19 is the

formulation of the viscous damping term. Aalbers does not include a wave induced water

particle velocity term in his development of the relative velocity. This should be a minor

component for the deep draft spar but is included in equation 3.18 for completeness.

Aalbers was not considering a deep draft vessel in his derivation and so the lack of this

component is puzzling.

Equation 3.18 is simplified to equation 3.20 by using the coefficients defined


Eq. 3.21

total linear (potential) damping coefficient:

spar linear (potential) damping coefficient:

wave linear (potential) damping coefficient:

non-linear (viscous) damping coefficient:

hydrostatic restoring force coefficient:

CL= b -= B +Bs
PA = h(1+ = + B


B b,
B Azh(,1+cr)

CN = b,

h(l +c.)

non-linear term coefficient: 1
h(l +Cmr)

spar added mass ratio: cm2
1 + cmr

wave added mass ratio: cm
1+ Cmr

The experimental testing was conducted without waves and with forced motion of

the spar. Removing the wave forcing from equation 3.20 results in equation 3.21 which is

the equation to be used to analyze the experimental data.

a +(CL +CN h -S)a +Kha +E 1-Lj h2 -h2 s(B C -- s.

Eq. 3.22

Analytic Approximation to the Complete Equation for Heave

Equations 3.20 and 3.21 lend themselves quite nicely to numerical solution since

they may readily be separated into a system of first order coupled differential equations.

They may also be coupled with similar equations for the motion of the spar itself.

However, it is desirable both from a computational efficiency standpoint and from

an application standpoint to obtain an analytic solution to the equations if possible. Such a

solution would allow the direct application of the resulting spar Response Amplitude

Operator function to a wave spectrum in order to obtain the response spectrum of the

system in random waves.

When the spar displacement (s) is small compared to the spar draft (So), as is

undoubtedly the case for the deep draft Chevron Spar, the term h= ha -s+SO may be

approximated by h So. Several methods of converting the non-linear damping to an

effective linearized damping were discussed in Chapter 2. These methods are usually

applied to less complex flows than those in this problem. However, an approximation to

the viscous damping in the form of that discussed by Clauss et al. may be appropriate.

Using the approximation discussed by Clauss et al. resulted in the effective damping

coefficient given in equation 3.22.

beff = b -s- ek= b,, c H cos- A cosO-e' Eq. 3.23

where HAMP, SAMP and 0 are the amplitudes and phases of the motion of the absolute

centerwell level and the spar, respectively, and 77o is the amplitude of the wave which is

assumed to have a cosine form. This approximation is discussed in greater detail in

Appendix A.

In their discussion of the linearization of the viscous damping, Clauss et al. noted

that, when linearized, this force "consists of a linear component with the same frequency

(as the wave), and an additional term with triple the frequency" (Clauss et al., 1992,

p.237). Equation 3.22 does not account for this triple frequency effect and thus a small

discrepancy is expected near one third of the synchronous frequency in the comparison

between a numerical solution of the equation of motion and this analytic approximation.

Utilizing the further approximation that z -So when the draft (So) is large

compared to the structure displacement (s) and the wave amplitude (1ro), the damping

terms can be combined into a single effective linear term as shown in equation 3.23 below.

h Ca +Cea +Kha +E 1l-Jh2 lS+Be,+ekz(p,ij+B,++K7y)
AR 2
Eq. 3.24

C, = C, + C, 8 -w I co )- S cos(0)-ekz

8 H S
B, = B, + C, mo co -^ cos0)-e
Be = BS + C 8 co 77, H C4) C4OS() ekI
31 77o, 71

Bwe = BW + CN, 8oo H co o) -S cos(0) -e .

Note that the 8c97, /37r parameter is that of Clauss, Patel and Berteaux, however

an equally valid parameter of 2or- 2/-7r from Ochi could be used.

Equation 3.23 still contains the non-linear terms associated with the E coefficient.

Examination of the relative magnitude of the terms in equation 3.23 may afford some

insight into the method to be used. Specifically, recognizing that E is inversely

proportional to the spar draft suggests that the associated terms may be neglected or

approximated by a perturbation technique with E as the perturbation parameter.

The first step is to non-dimensionalize equation 3.23 by recognizing that all terms

have dimensions of acceleration and then dividing by the gravitational acceleration (g).

Assuming that the solution is sinusoidal, Table 3.2 shows the orders of values (where

appropriate) and the orders of magnitude of the components of equation 3.23.

Table 3.2, Order of Magnitude of Individual Terms in Heave Equation

Term (s) Approx. Value Order of Magnitude

ha, a, a, s, s, ql, ~1-10 (m, m/s, m/s2) 1 10

g 9.8 m/s2 10

h, So -200 m 100

S-1 0 to -0.5 0.1

Ai -500 m2 100

Cm, cM, cM -0.1-1 0.1-1

p 1025 kg/m3 1000

e-kSo 0.002 (for T = 12 sec) 0.001

Ce, Bse, Bwe (see below) -0.1 1

Assume that the damping coefficient takes the form of b = pCfA,, where Cf is

the coefficient of friction and As is the surface area. Consider As to be the surface of the

interior wall of the centerwell so A.s = rDSo (recall Di is the centerwell cross-sectional

diameter) and Az = nD2 From Hoerner (1965, p. 3-11 to 3-13), Cf can be estimated to

be of order 0.1 to 1. The result is the following equation for the damping coefficients of

equation 3.23.

b pCfDh C, 0(0.1)
pAh(l+ ) D +c 2 (l(0.0+c1)(10)0(1)
pAh(1+c,, pPr-Dh(1+ cl ) D (1 + c.) O(10)O(1)

Table 3.3, Orders of Magnitude of Complete Terms

ha g __ o(1)
,g' 0() = 0(0.1)
g gg 0(10)
CL BS Bw C B, B 0(0.01) or 0(1) ( o
g g g g' g g 0(10)
K 1 0(1)- 0(0
g h(1+ cm) 0(102)0(1)

E 1 o0() = 0(10-
g h(1+ c,) 0(10)0(102)0(1)

(1- _L)2 2=
AR. 2 I O(0.1)0(1) 0(0.1)
( -R)ha 2(1- )sha +(1- AR2

s ,w cm. 0(1) -0(0.1)
g 'g g(1 + c) 0(10)0(1)

F- (W e (i + Bi + Kr ) (0(0.1) + O(0.1) + 0(0.1)) 0(10-5)
g g 0(10)

If the relationships of Aalbers (1984) are used, the order of magnitude of the

prototype size damping is 0(0.01)0(D )=0(0.01)0(100)= 0(1). Applying the

estimates for individual terms in Table 3.2 to the non-dimensionalized terms in equation

3.23, results in the orders of magnitude for the complete terms in Table 3.3. The wave

force terms have been combined into a single term Fw which has units of force per unit


Thus, the order of magnitude of the non-dimensionalized terms of equation 3.23

may be summarized below.

h Ch Kh E 1[rj 1 1 Bs F
_a + e + a2 + iL 2 + _e + w
g g g g AR 2 g g g
0(0.1) 0(0.1) 0(0.01) O(10-4) 0(0.1) 0(0.1) O(10-5)

Therefore, it is readily apparent that both of the terms associated with the E

coefficient and the wave forcing are small compared to the rest of the terms. The wave

forcing terms are linear and may be incorporated in the first order equation. The other

small term may be treated by the perturbation technique described by Ochi (1990, p. 398),

with the perturbation parameter s=E= 1/S(1 +c,). Letting ho be the first order

solution and hi be the second order solution, equation 3.23 may be separated into the

equations below.

(a) h +C,ho +Kho = /ss+B,S+F, and
Y [ >\^* \2 1 2 Eq. 3.25
(b) h, +Ceh1 +Kh, =- 1- Eq. 3.25-h .

IfFw, s, ho, and hi are assumed to be sinusoidal, equation 3.24 (a) will result in a

solution of the form h =f, (Ao cos cot+Bo sin cot) where fa is the forcing function

amplitude. Substituting this into equation 3.24 (b) and manipulating the squared terms

will result in a solution of the form h1 = fa (A1 cos2ct +B, sin2cot +Y). The final form

of the solution is given in equation 3.25. Note that this equation predicts a mean

centerwell surface level displacement (fas Y) which results from a reduction in pressure at

the bottom of the centerwell due to flow through the orifice.

ha = f [(A0 coscot + Bo sin cot) + e(A1 cos2cot + B1 sin 2ot) + 6Y] Eq. 3.26

A response amplitude operator (RAO), a frequency response function (H(co)) and

the lead angle (b) can then be determined. The complete analytic solution to equation

3.23 and the equations forAo, Bo, A and Bi is given in Appendix A.

To summarize, the approximations made in obtaining an analytic solution to

equation 3.20 are: (1) the wave amplitude and the amplitude for the spar and centerwell

water surface motion are small compared to the spar draft so that h = SO and the position

of the bottom of the spar is approximately -So, (2) the non-linear damping term may be

approximated in the manner of equation 3.23 to yield an effective linear damping term, (3)

the terms associated with the E coefficient of equation 3.20 may be assumed to be small

for large spar drafts and treated by the perturbation technique with the perturbation

parameter e = E. The determining assumption in these approximations is that the spar in

question is deep drafted. Thus, the solution developed should not apply to the moonpool

experiments of Aalbers, however even Aalbers ultimately linearized his equation by simply

neglecting the non-linear terms associated with the E term of equation 3.20. The final test

will, of course, be the ability of the solution to correctly reproduce the experimental

results to be described in the next chapter.


Model Description and Scaling

Experiments were conducted in a wave tank at the University of Florida Coastal

and Oceanographic Engineering Laboratory. The water portion of the tank is 1.2 meters

(4 feet) deep, 0.6 meters (2 feet) wide and 28 meters (92 feet) long. The sides are

constructed of 1.2 meter (4 foot) square glass panels.

The model was constructed of concentric transparent Plexiglas tubes capped at the

end except for the centerwell area. The transparent model in conjunction with the

transparent side panels of the wave tank allowed visual observation of the centerwell

surface level displacement. Such an observation was vital to locating and examining the

synchronous motion noted in Chapter 1.

Following the lead of Knott and Flower (1980, p. 164), it was decided to maintain

geometric similarity as much as possible between the model and the Chevron Spar

prototype. The size of the available tanks for conducting the experiment and the desire to

minimize bottom influence restricted the model draft to approximately 0.6 meters (2 feet).

To maintain geometrical similarity, the outer diameter was approximately 13 centimeters

(5 inches) and the centerwell diameter 7.6 centimeters (3 inches). These dimensions yield

diameter-to-draft ratios of 0.22 and 0.13, respectively, which are those of the Chevron

Spar design. The resulting length scale is 1/325 or 0.003

The standard procedure in ship modeling is to maintain Froude scaling between the

model and the prototype. The total resistance is determined based on model tests and the

skin friction is calculated mathematically based on Reynolds number. The difference

between the skin friction and the total resistance is called residual resistance and is

assumed to be a function of the Froude number only. Thus the residual resistance may be

scaled directly from the model tests and the total prototype resistance is found by adding

the scaled residual resistance and the skin friction calculated for the prototype Reynolds

number (Todd, 1967, pp. 291-293; Randall, 1997, p. 159).

However, this leads to a problem for the oscillating flow of the spar platform in

waves since the Reynolds number varies, making the skin friction coefficient variable with

time. For the spar itself, many authors neglect the effects of viscous (Reynolds number

dependent) forces all together, thus eliminating the problem. (Kokkinowrachos and

Wilckens, 1974, p. 99). Unlike the solid spar, the motion of the fluid in the centerwell

may be influenced by these viscous friction forces. Through physical experiments, Aalbers

found that quadratic damping is proportional to the diameter of the moonpool squared

(Aalbers, 1984, p. 561). Thus, Froude scaling may be used in the experiments and the

scaling of friction forces to the prototype size should be adequately approximated by

Aalbers relationship.

As constructed, the model's total length (draft plus freeboard) was 76 centimeters,

the outer diameter, Do, was 14 centimeters and the inner (centerwell) diameter, DI, was 8

centimeters giving a wall thickness of 3 centimeters. The mass of the model was

augmented with lead weights at the bottom and inside the model wall. This gave the

model an overall mass of 6.6 kilograms (14.5 pounds) and a still water draft of 64.5

centimeters. The model-to-prototype scaling factors thus obtained were 1/310 for the

length scale and JL./L, = =1/310 = 1/17.6 for the time scale (Froude scaling).

Three orifice plates were constructed to install on the bottom of the spar and

partially obstruct flow through the centerwell. Orifice size was selected to provide as

significant an obstruction as possible without exceeding the arbitrary limit of 50 percent

area reduction discussed in Chapter 3. Table 4.1 summarizes the orifice diameter (DR) and

orifice area ratio (AR/Ai) information and Figure 4.4 illustrates the model and orifices to


Table 4.1, Orifice Plate Diameters
Orifice Plate No. Orifice Diameter (DR) DI/D A AI= AR
1 6.9 centimeters 0.86 0.74
2 6.4 centimeters 0.8 0.64
3 5.8 centimeters 0.72 0.52

Free Response Tests

Free response experiments were conducted by displacing and releasing the model

in still water. The tests were conducted first with the centerwell fully open (referred to

here as the "open bottom" configuration) and then with the centerwell completely closed

(the "closed bottom" configuration). In the closed bottom configuration, the centerwell

was filled with water to maintain the same draft as the open bottom configuration.

Do= 14cm
DI= 8 cm
DRI= 6.9 cm
DR2= 6.4 cm
DR3= 5.8 cm






Orifice 1
DR/ D = 0.86
AR/ A = 0.74

Orifice 2
DR/D, = 0.8
AR/ A = 0.64

Orifice 3
SDR/ D = 0.72
AR/ A = 0.52

76 cm

^ '

Fig. 4.1, Scale drawing of model and orifice plates.

A video recording was made of the model's motion against a grid placed on the

tank side panel with a line on the model to mark the still waterline. Following the

experiments, the video recording was reviewed frame by frame and the waterline mark

noted as it passed each grid line. Special note was made when the spar crossed the zero

level and at each peak and trough of its motion.

The system was assumed to be linear and the damped natural frequency (fd in Hz,

COd in radians/sec) and the logarithmic decrement (6) were calculated from the video data.

Once these values were obtained, equation 4.1 and the definitions below were used to

D, Do

v~f v

calculate the added mass coefficient cm = a/m and the linear damping ratio

= b/2J(m+a)k Table 4.2 shows the results of these calculations for both model


m(l + c, ) + bs + ks = 0
6 8k k Eq. 4.1
P=-, =o- a=----m
2n m+a) C

Table 4.2, Summary of Free Response Damped Natural Frequency (fd),
Added Mass Coefficient (c,,) and Damping Ratio ()
Run Configuration fa (Hz) c, f8
1 open bottom 0.614 0.023 0.035
2 open bottom 0.610 0.037 0.032
3 open bottom 0.606 0.050 0.039
16 closed bottom 0.600 0.073 0.021
17 closed bottom 0.598 0.080 0.022

A noticeable difference exists between the added mass and damping ratio for the two

configurations. The added mass coefficient of the model increased when the bottom is

closed off indicting that the more bluff the body the greater the added mass. The damping

ratio increases when the bottom is opened, presumably due to the increase in surface area

subject to friction and the vortex generation by the bottom edge of the centerwell. Both

of these results were expected.

The value of the damping ratio was of the order of magnitude (-0(0.01)) expected.

The added mass coefficient is comparable to that of an ellipsoid with a large major axis to

minor axis ratio and flow parallel to its major axis (Liggett, 1994, p. 114). A worthwhile

comparison would be between the added mass obtained in the experiments and that

computed using the method of Clauss et al. (1992, p.281 and shown in Figure 2.3). This

calculation results in added mass coefficients of 0.082 and 0.098 for the open bottom and

closed bottom configurations, respectively. Using the two computed added mass

coefficients, damped natural frequencies (fd) of 0.596 Hz and 0.592 Hz, respectively,

were calculated. The coefficients are relatively close to the values calculated from the free

response test data and the damped natural frequencies are certainly within the error of the

experimental data.

Only a few of the experiments produced useable data. The rest were adversely

affected by horizontal model motion disrupting the video focus, poor visibility of the

waterline mark, the model pitch motion, and other problems associated with the

unrestricted motion of the model. Despite the problems encountered with the free

oscillation experiments and the measurement techniques, the experiments did provide

useful information about the behavior of the spar with a centerwell. The possibility of

using the added mass coefficients of Clauss et al. was confirmed. Also, the damping ratio

was found to be small in both configurations.

Configuration for the Forced Motion Experiments

The bulk of the experimentation was conducted by directly forcing spar motion

and measuring the response of the centerwell water level with electronic instrumentation.

The instrument output was recorded using an analog-to-digital conversion card and Global

Lab data collection software. The sampling frequency was 60 Hz with the run times

varying from 2 to 4 minutes. The data were analyzed using programs written in MATLAB

which allowed for interactive graphical examination of the data and the use of built-in

signal processing functions to analyze the time varying experimental results.

To conduct the experiments, a frame was constructed to house the model, provide

a platform for the drive mechanism and instrumentation and to restrict the motion of the

spar to heave alone. The frame consisted of a square upper and lower plate 46

centimeters (18 inches) on a side. The plates were separated by 122 centimeter (48 inch)

rods at each corner. The frame stood on the lower plate and the model was located inside

at an elevation which provided approximately 36 centimeters of separation between the

bottom of the model and the lower plate at the lowest point in the model's stroke.

Attached to each plate and running through special tubes in the wall of the model were

four guide wires designed to minimize twisting and horizontal movement of the model.

Attached to the top of the model was a collar. Through the collar, on either side

of the model, were two drive rods which penetrated the upper frame plate through slide

bearings. Above the upper plate, the drive rods went through, but were not attached to a

lower yoke. The drive rods were then connected to an upper yoke. A strain gage was

located between the upper and lower yokes. A third drive rod was attached to the bottom

of the lower yoke, directly below the strain gage. This drive rod was connected to a fly

wheel on its lower end. The fly wheel was attached to a variable speed, rotary AC motor

which provided the driving force for the tests. The apparatus is illustrated in Figure 4.2.

The instrumentation consisted of the strain gage mentioned above; a position

indicator attached to the upper yoke to measure the spar displacement; a capacitance type

wave gage inside the centerwell for measuring the centerwell water level; and a

capacitance wave gage located outside of the system.

The design of the position indicator was based on the capacitance wave gage. This

device consisted of a stationary metal tube covered with rubber inside a second metal tube

which was free to move over the stationary tube. The stationary tube was mounted to the

top of the upper plate of the model frame and the moving tube was attached to the upper

yoke of the drive assembly. As the yoke moved up and down, the moving tube exposed

varying amounts of the stationary tube causing the capacitance to change. This changing

capacitance was measured to give the position of the spar. This instrument, which was

designed and constructed at the Coastal Engineering Laboratory, proved to be a reliable

and stable position indicator with a linear response.

The motor initially used to drive the spar could not provide periods longer than

approximately 1.5 seconds and motion in vicinity of this period was uneven and strained.

The ideal natural period of the centerwell (i.e. the natural period without accounting for

the damping or added mass) is 1.6 seconds ( T = 2r / g/So ), hence the period of

interest for investigating the synchronous motion of the centerwell was longer than the

upper limit of the driving mechanism. The result was that the synchronous motion was

not witnessed during the initial testing due to the under-powered motor.

In addition to the problem of an under-powered motor, the drive linkage proved

unstable. The circular motion of the motor shaft was not well translated to the desired

linear heave motion and significant lateral motion of the drive rods was produced. While

the upper plate drive rod bearings and the guide wires limited the motion of the model to


the heave direction, the oscillation of the drive mechanism resulted in the entire frame

vibrating. Thus, the resulting centerwell response was adversely influenced by secondary

vibrations of the frame inducing seiching in the centerwell.

Fig. 4.2, Frame and drive assembly.

The first motor was replaced by a larger motor capable of driving the model at

longer periods. To minimize the lateral oscillations of the drive rods, the upper plate

bearings were reinforced and the frame was weighted. These adjustments solved most of

the initial difficulties, however the reinforced bearings produced excessive friction forces.

The magnitude of the friction forces was such that the damping and added mass could not

be extracted from the force data.

_ position

upper yoke 1---

lower yokeposition indicator Drive Assembly
ler ye support assembly
drive rod
d drive rod bearings

upper frame plate- 10 cm position
collar i indicator
upper yoke

lower yoke -
122 cm
model --l w ii-
drive rod
-- bearing
loe fm 36 cm
lower frame plate -


fly wheel

However, as previously noted, the free response tests indicated the method of

Clauss et al. provided an adequate added mass estimate for the Spar and numerical means

of estimating the damping coefficient were available (see Chapter 2). Thus the loss of

these data was not considered significant to the overall accomplishment of this study.

Experimental Results

A total of 88 separate experiments were conducted. The experiments covered

oscillation periods from 0.6 seconds to 6 seconds which correspond to prototype wave

periods from 10.6 to 105.6 seconds. Shorter periods were not tested due to continued

vibrations experienced by the frame at higher frequencies. In addition, as noted in Chapter

3, the long wave approximation becomes suspect at periods shorter than 12 seconds

(prototype scale). The majority of the experiments were performed with a constant

amplitude of approximately 2 centimeters, although some early tests used a 4 centimeter

stroke. Analysis of the data showed that the stroke difference did not affect the results.

Notable was the absence of radiated waves during the experiments. Radiated

waves where not observed or measured by the wave gage located outside of the system

(except within the first centimeter of the model). Even the small response in the

immediate vicinity of the model was minor, on the order of millimeters or less.

As previously mentioned, the model position and centerwell level displacement

data was analyzed using MATLAB software. Since the centerwell wave gage was

mounted to the model itself, the resulting water level information was relative to the

model position. The information of interest was the position of the centerwell level with

respect to a fixed coordinate system. This was obtained by the simple formula

ha = h + s- S; where ha is the centerwell level with respect to a fixed coordinate system

(absolute centerwell level), h is the height of the water column inside the centerwell with

respect to the base of the model and s is the model displacement (see Figure 3.1). Thus

the data to be analyzed was itself derived data and not directly measured.

The data were analyzed using two methods. After deriving the absolute centerwell

level (ha) a Fourier analysis was conducted using the Fast Fourier Transform in MATLAB

to obtain frequency, amplitude and phase information for the relative and absolute

centerwell levels and the model position. Phase was measured as a lead angle with respect

to the model displacement. The average frequency bin width and therefore the accuracy of

the frequency was 0.03 Hz. For the most part, the spectrum of the relative centerwell

level and the model displacement gave solid single frequency responses. In all cases where

multiple frequencies existed, the secondary frequencies were within the neighboring bin or

the magnitude of the secondary response was several orders of magnitude less than the

primary response.

This was not the case for the absolute centerwell levels. Though the majority of

the tests showed the same response as above, five cases gave very different results. In all

of these cases the frequency was high and the amplitude response was small indicating a

decoupling of the centerwell motion and the model motion. This could be expected since

four of the five cases were tests with no orifice where the coupling is weak. In the case

where this occurred with an orifice (AR = 0.64), the frequency was near the top of the

experiment band (1.39 Hz, o/Oo = 2.2) and the response was only 2 percent of the model


The Fourier analysis yielded good response amplitude results, but the phase results

left much to be desired. When the absolute centerwell displacement was simulated using

the frequency, amplitude and phase information obtained by the Fourier analysis and the

time series was compared to the data time series, the results were good for amplitude and

frequency matching, but inconsistent for the phase. Also discernible from examination of

plots of the data with time were variations in the speed of the motor itself. Even the

model displacement times series, which was the most consistent of the three sets, exhibited

a small frequency variations due to instabilities in the motor speed.

The hope for correcting this situation lay in the singularity of the spectra. The

spectra showed a concentrated spike exclusively at one frequency. This suggested that the

data could be analyzed by locating the local maxima and minima for each time series.

With the time and magnitude of the local maxima and minima determined, mean values for

the frequency, amplitude and phase of the time series were calculated. This method is

referred to from this point on as the "maxima-minima analysis."

The maxima-minima method produced excellent results. The information

developed for frequency, amplitude and phase were evaluated by comparing a generated

time series with the experimental data and qualitatively ranking the results as good, fair,

poor or unsatisfactory. A rank of good was given to 46 of 88 of the data set comparisons

(52 percent), 24 ranked fair (27 percent), 13 were poor but usable (15 percent) and only 5

were deemed unsatisfactory and unusable (6 percent).

A product of the maxima-minima method was the variance of the means of

frequency, amplitude and phase which could be used as an indication of the experimental

error. The frequency variance yielded a maximum standard deviation of 0.03 Hz which

was consistent with the frequency bin width of the Fourier analysis. Drift in the position

and level instrumentation gave maximum amplitude errors of 0.06 centimeters and 0.15

centimeters for the model displacement and relative centerwell level, respectively. This

results in an error of 0.21 centimeters for the derived absolute centerwell level. The

variance of the mean amplitude from the maxima-minima analysis showed average

standard deviations of 0.02 centimeters, 0.14 centimeters and 0.12 centimeters for the

model displacement, relative and absolute centerwell level, respectively. Therefore, error

for the model displacement, relative and absolute centerwell levels was set at the average

of these; 0.04 centimeters, 0.15 centimeters and 0.17, respectively. To examine the ratio

of the amplitude of the absolute centerwell level and the model displacement amplitude,

the errors above were converted to percentages of their respective measurements. This

gave the maximum errors for the model displacement and absolute centerwell level as 1

percent and 9 percent, respectively, thus value for the error in the amplitude ratio was

approximately 10 percent.

The phase error estimate was more difficult to calculate. The phase of the

centerwell level motion was calculated with respect to the phase of the spar. However,

since both the period of the centerwell motion and the spar were used to calculate the

phase, their respective errors must be included in the phase error estimate. With a

frequency error of 0.03 Hz (for each frequency in the calculation) and a maximum period

of 2.5 seconds, the phase error may be estimated by AO = 2(1/Af / T,,) = 27 degrees.

Table 4.3 contains a summary of the experimental error values.

Figures 4.4 through 4.7 present the experimental results as plots of the absolute

centerwell level to model displacement amplitude ratio and phase angle difference between

the absolute centerwell level and the model displacement as functions of the frequency

ratio. The phase angle is a lead angle such that if the motion of the model is of the form

cos(x), the absolute centerwell level response is cos(x + 4) with 4 the lead angle

represented in the plots. The natural frequency of the centerwell was expected to be

dependent on the orifice size. Therefore, when comparing the orifice configurations, the

frequency ratio for the plots is defined as the response frequency (o)) divided by the ideal

natural frequency (ca,) given by cw, = g/So From this point on in this thesis, the

natural frequency of the given system will be referred to as o_ or V-K (see equation 3.21).

Note that this frequency is also the "synchronous frequency."

Both the results of the maxima-minima and the Fourier analysis are presented. As

previously noted, the amplitude ratio was consistent between the two methods, but the

phase angle shows a decided shift for the data analyzed with Fourier analysis.

Reflected in the plots is the distinct peak in the amplitude ratio corresponding to

the point where the centerwell level moves in sync with the model. In other words, the

water in the centerwell moved as though the bottom of the centerwell was closed at the

bottom. The synchronous response phenomena occurred in all cases except the no orifice

case shown in Figure 4.7.

Table 4.3, Estimated experimental error values.
Data Estimated Error
frequency 0.03 Hz
model displacement (s) 0.04 centimeters
relative centerwell level (h) 0.15 centimeters
absolute centerwell level (ha) 0.17 centimeters
amplitude ratio (hls) 10 percent
phase angle 27 degrees

For the data developed using the maxima-minima analysis, the phase plots show a

zero phase angle at the frequency corresponding to the synchronous motion. However,

the data developed from the Fourier analysis show a 30 to 40 degree lead angle for this

point in all of the plots. Since observation confirmed the zero phase angle, the Fourier

analysis seemed to have introduced some bias in the data. The maxima-minima analysis

results indicated the observed zero phase angle. All comparisons made beyond this point

are based on this data extracted by the maxima-minima analysis.

It was noted that the no orifice condition did not produce synchronous centerwell

motion. Observation showed that the synchronous motion could be initiated without the

orifice, but could not be maintained for any significant length of time. When the time

series were processed, the short periods of synchronous motion were averaged out by

both analysis methods.

The no orifice configuration has a significant response over an extremely narrow

frequency range. The amplitude ratio for this configuration drops very rapidly as

frequency moves away from the natural frequency of the system. As the orifice size

decreases, the slope of this drop grows shallower and the frequency range increases. As

expected, this suggests a weak coupling between the centerwell and the model over most

of the frequency domain when the orifice is not installed. This is also reflected in the great

variability of the phase data as compared to the configurations with the orifice installed.

Examining the data, and especially the phase data, it is apparent that the overall

centerwell response becomes more consistent as the orifice size decreases. In other

words, the coupling between the centerwell fluid and the model itself is increasing with

decreasing orifice diameter. This is expected since it is primarily the orifice which

maintains the coupling between the centerwell and the model. This coupling may be

attributed to the model acceleration term (ju, ) of equation 3.22 and the damping and

induced vortex formation due to the sharp edge at the centerwell exit and at the orifice lip.

Also, the non-linear terms associated with the E coefficient have a dependency on the

orifice area and so have an influence on the coupling.

The lack of success in maintaining the synchronous motion in the no orifice case

does not necessarily imply that such motion is not possible for this configuration. The

absence of such a response in these tests may be due to inadequate control of the model

oscillation frequency combined with a narrow frequency band for attaining synchronous

motion. The very narrow frequency band over which synchronous motion occurs in the no

orifice case makes this condition difficult to achieve. Slight shifts in the motor frequency

as noted in the data analysis could easily cause the response to move into the regions

where little or no response exists.

The frequencies at which synchronous motion occurred were estimated from the

data plots and are presented in Table 4.4. The table lists the orifice diameter ratio (DR)

and orifice area ratio (AR) and the corresponding frequency, radian frequency and

frequency ratio. Figure 4.3 displays the frequency ratio versus the orifice diameter ratio

and includes a linear least squares fit line to demonstrate the excellent consistency of the

data. In Chapter 4 this information will be used to evaluate the centerwell added mass


Finally, referring to only the data developed by the maxima-minima analysis, there

is a distinct dip in the phase angle for frequency ratios of less than 0.5. This may be

attributed to the 2o term in the analytic solution described by equation 3.26.

Table 4.4, Synchronous motion frequencies (an) by orifice size.
DR AR frequency (Hz) o. (radians/sec) On/co
0.72 0.52 0.5974 3.7536 0.9629
0.80 0.64 0.5940 3.7322 0.9574
0.86 0.74 0.5922 3.7209 0.9545
1.0 1.0 0.5844 3.6719 0.9420

Synchronous Frequency Variation with Orifice Dia. Ratio




8 y =-0.0749x+ 1.0175
0.95 R2 = 0.9877


0.94 -
0.70 0.75 0.80 0.85 0.90 0.95 1.00

Fig. 4.3, Plot of synchronous frequency with orifice diameter ratio.

Amplitude Ratio Plot for Orifice Area Ratio = 0.52
+ maxima-minima X fourier



S0.6 +

^ +
| 0.4 x
x +
0.2 x + *'
W + +
Sx "+ x
0.0 0.5 1.0 1.5 2.0 2.5

Phase Plot for Orifice Area Ratio = 0.52

+ maxima-minima x fourier
xx x

30+ +

0 t x

30 x
" +

-60- +
| -30 ---------T---- ------

+ + + + +

-90 + x
0.0 0.5 1.0 1.5 2.0 2.5

Fig. 4.4, Results for experiments conducted with an orifice area ratio of 0.52

Amplitude Ratio Plot for Orifice Area Ratio = 0.64

+ maxima-minima x fourier


0! X
. 0.8 x


0.0 0.5 1.0 1.5 2.0 2.5
O /0 o

Phase Plot for Orifice Area Ratio= 0.64
+ maxima-minima X fourier
0.x x


-60 +

0.0 0.5 1.0 1.5 2.0 2.5

Fig. 4.5, Results for experiments conducted with an orifice area ratio of 0.64.
60 -------+5 --------------

-60 ----------- x -------- x

-90 ---------------------

Amplitude Ratio Plot for Orifice Area Ratio = 0.74
+ maxima-minima x fourier


" 0.8-

" 0.6

g 0.4
ss x

xx ++ +

0.0 0.5 1.0 1.5 2.0 2.5

Phase Plot for Orifice Area Ratio = 0.74
+ maxima-minima X fourier
+ +
+ + + x+4
60 --PX--
F X +
C 30 +--
'^ 0 --- ----,--------------

S-30 ----X

-60 x
+ + +x

0.0 0.5 1.0 1.5 2.0 2.5
0o/0 o

Fig. 4.6, Results for experiments conducted with an orifice area ratio of 0.74.

Amplitude Ratio Plot for the No Orifice Case
+ maxima-minima X fourier


S0.8X +



+ x +x +

0.0 1---+-----+-------------+----
0.0 0.5 1.0 1.5 2.0 2.5 3.0

O/CO o

Phase Plot for the No Orifice Case
+ maxima-minima x fourier

+ + ++
30 --
M)O o---- y--------
x x x
S-30 --

X X. + x
S-90 +
-120 -

-150 --
0.0 0.5 1.0 1.5 2.0 2.5 3.0

Fig. 4.7, Results for orifice experiments conducted with an orifice area ratio of 1.0 (i.e. no


Model Calibration

The equation of motion for the centerwell driven by induced spar motion without

waves was expressed in equation 3.22 and is repeated below.

a, + (CL + C a + Kh + EL(1- 2 = + (Bs + CE A s)

With the spar undergoing forced motion, the first assumption is that the linear

damping terms are equal, that is CL = Bs. This seems reasonable since both terms

describe the same interaction between the spar and centerwell. This results in four

unknown parameters to be determined from the data described in Chapter 4; the linear and

non-linear damping coefficients, CL and CN the added mass coefficient associated with

the spar acceleration and orifice lip, Cm2, and the added mass coefficient associated with

the centerwell motion, Cm. If cm2 is assumed to be small compared to Cm than the terms of

equation 3.22 involving the added mass may be approximated as follows:

s = Cm2/1+ ,, K= g/[So(1 + cm)] and E= 1/[So (1 + c)]).

Note also from Chapter 3 that the added mass coefficient (cm) is embedded in the

damping coefficients as 1/( + Cm)

When the centerwell is experiencing synchronous motion, there is no relative

velocity between the centerwell and the spar. Thus, for this case (and with CL = Bs),

equation 3.22 becomes equation 5.1 below,

ha +Kha ,- E == Pus. Eq. 5.1

In the synchronous motion condition the centerwell acceleration must also match

the spar acceleration and 'ha ug = (1- jsI)ha. The E term was shown to be small in

Chapter 3 and is smaller still without its relative velocity component (h2). Thus, equation

5.1 can be further simplified to equation 5.2 below.

h, + Kha = 0. Eq. 5.2

Equation 5.2 is the equation of motion for a mass-spring system undergoing a

free, undamped oscillation. The fact that the centerwell in synchronous motion is

equivalent to such a system is an interesting result in itself and allows the added mass

coefficient to be calculated from the synchronous motion frequency, cO, by

c, So /gco2 1. Table 5.1 lists the added mass coefficient thus determined by orifice

diameter ratio and area ratio.

Table 5.1, Added mass coefficient derived from the
synchronous frequency of the centerwell
DR AR On, (radians/sec) Cm
0.72 0.52 3.754 0.078
0.80 0.64 3.732 0.091
0.86 0.74 3.721 0.097
1.0 1.0 3.672 0.127

Aalbers found that the added mass of the of the moonpool was proportional to the

radius of the moonpool to the third power. Using this relationship, the following can be


i D3 D
a = cm = c 4D2So oc D' -- cm oc --
m4 D2So So

Since the synchronous frequency varied with orifice diameter, it was concluded

that the diameter of interest is the orifice diameter. Plotting the added mass coefficients of

Table 5.1 against the orifice diameter to spar draft ratio results in Figure 5.1. The plot

shows a linear variation of the added mass coefficient with the orifice diameter to spar

draft ratio. Equation 5.3 expresses this relationship.

m 1.4 -1-0.05 Eq. 5.3

Plot of Added Mass Coefficient vs. DR/So

I 0.12
0.11 -
0 0.09-
0.08 y = 1.3993x 0.0483
S0.07 R = 0.9859
0.06 I
0.08 0.09 0.10 0.11 0.12 0.13

Fig. 5.1, Added mass coefficient vs. the cube of the orifice diameter ratio

Next consider the non-linear damping coefficient. The model is essentially a pipe

with fluid flowing through it, so the methods used to calculate flow through pipes should

be applicable. The Darcy equation, which is used to calculate head loss in a pipe, is given

in equation 5.4:

hL =k Eq. 5.4

where hL is the head loss, v is the velocity of the fluid and k is the resistance coefficient

(Crane, 1976, p. 2-8).

The non-linear damping coefficient can be divided into components derived from

pipe flow processes and represented by different resistance coefficients. These coefficients

when summed should yield the non-linear damping coefficient, Cs.

The resistance coefficient for the pipe itself is k = fL/D where is the friction

factor and L/D is the length of pipe compared to its diameter. Multiplying the head loss by

the hydrostatic coefficient (K= g/So(1I+ c.)) will convert the form of equation 5.4 to

that of equation 3.22, i.e.

g- hL =( ) (f L 1 Iv vI. Eq. 5.5
So( +cm) D 2S2(1+c) + 2S(l1+cm)

The model also has an abrupt entrance and exit. The resistance coefficients for an

abrupt entrance and exit are 0.5 and 1, respectively (Crane, 1976, p. 2-11). Since the flow

is oscillatory, the bottom of the centerwell will act as an entrance over half the cycle and

an exit over the other half. Thus, averaging these resistance coefficients and applying the

conversion of equation 5.5 gives the component of the non-linear damping associated with

the entrance and exit losses. The resistance coefficient for this component is defined by

equation 5.6.

(ken k vv 0.5 1 vIV 0.375
[ +entrance +. - +- = v v Eq. 5.6
2 2 2So(1+cm) 2 2 2So(1 +c) So(1 +c .)

Finally, the last component of the non-linear damping to be considered is the effect

of the orifice itself. From Crane (1976, p. 3-5) the volumetric flow rate, q, through an

orifice in a pipe can be described by equation 5.7:

CAR 2 gh-
q =-R Eq. 5.7
1 DR4

where AR is the orifice area, DR is the diameter of the orifice divided by the diameter of

the pipe and C is a flow coefficient.

Converting the volumetric flow rate in equation 5.7 to the flow rate in the

centerwell calculated from the absolute centerwell level velocity (a) and squaring both

sides leads to equation 5.8, with C' = 1/C2 and DR4 = AR2.

A CAR 2gh hA h C__
l-DR4 AR AR 1- AR2

2 C2 2ghz (1-AR2)
a Ih =. ->h
AR2 1-AR2 L 2gC2AR2

g hL~ C' 1 -l ha Eq. 5.8
So(1+cm) 2So(l+c) AR2 )

Summing the components in equations 5.5, 5.6 and 5.8 gives equation 5.9 which

expresses the complete non-linear damping term.

CN 0.75+ (f +C'R 2 1 Eq. 5.9

Note that the C' term disappears for the no orifice case when AR = 1. Neglecting

the linear damping term (CL) and tus, a least squares fit of the amplitude ratio data for the

no orifice configuration resulted in a value for fLID of 2.8.

Using this value, another least squares fit was performed on the data collected for

the configurations with the orifice installed. The C' coefficient was found to have a

frequency dependence which resulted in different values above and below the synchronous

frequency (ca,). The equations below summarize the least squares fit results for the non-

linear damping (with linear damping neglected).

(f = 2.8

C' = 2.65 for DR <0.8 and co < co,,

C' = 6.5 1-- DR for all D and co > co,,

Still neglecting the ,Us and the linear damping terms, the above relationships were

applied and the equation of motion was solved both numerically and analytically. The

numerical solution was achieved using an adaptive step controlled fifth order Runge-Kutta

algorithm with an embedded fourth order algorithm written in Fortran 90. A time series

was obtained and analyzed using the maxima-minima method. The analytical solution of

Chapter 3 and Appendix A was likewise computed using a program written in Fortran 90

and later rewritten in MATLAB.

Figures 5.2 and 5.3 display the results of these solutions for the configuration with

an area ratio of 0.74. Figure 5.2 depicts the ratio of the centerwell and spar motion

amplitudes and the phase difference between the centerwell and spar motion as functions

of frequency normalized to the ideal natural frequency of the system. The experimental

data for this configuration is also depicted. Figure 5.3 shows the mean centerwell level

offset noted in the analytical solution of the equations of motion found in Chapter 4. This

phenomena was not anticipated prior to performing the experiments and the effect cannot

be extracted from the data collected.

The numerical and analytical solutions show excellent agreement between one

another for the amplitude ratio and the mean centerwell level. Only the slightest of

disagreements occur at very low frequencies and near the synchronous frequency. Both

solutions show a small peak in amplitude at half the synchronous frequency due to the 2c0

term found in the analytical solution. The numerical solution gives a 30 peak as discussed

by Clauss et al. resulting from the non-linear damping term which was not accounted for

the linearization process for the equation solved analytically.

A large difference in the phase results of the solutions at low frequencies

corresponding to these 2o) and 30 effects and oscillations in the numerical solution phase

at higher frequencies illustrates the extreme sensitivity of the phase to secondary effects.

This explains the difficulty in extracting reliable phase information from the data discussed

in Chapter 4.

Non-Linear Damping Amp Ratio Plot for AR = 0.74
+ maxima-minima analytic -- numeric



| 0.6

+ s / + _____+
+ + +
0.0 0.5 1.0 1.5 2.0 2.5 3.0

Non-Linear Damping Phase Plot for AR = 0.74
+ maxima-minima analytic -- numeric

90 -


S30 +

P 0

g -30

+ +
-90 -

0.0 0.5 1.0 1.5 2.0 2.5 3.0
o /0 o

Fig. 5.2, Mathematical model solutions for area ratio 0.74, non-linear damping only with
Ps = Cm2 = 0.

Non-Linear Damping Mean CW Level Plot forAR = 0.74
analytic -- numeric

-2.E-04 -




0.0 0.5 1.0 1.5 2.0 2.5 3.0

Fig. 5.3, Mean centerwell level variation with frequency determined from numerical and
analytical solution of the governing equation.

On the whole, agreement with the experimental data is good. Similar results were

obtained for the experimental data with other orifices. The solution amplitude ratios do

not agree as well with the experimental data at lower frequencies, but the experimental

data itself is somewhat questionable at these frequencies suggesting scatter due to

additional secondary effects unrelated to the heave motion. Possible causes of such

secondary effects are centerwell seiching and drive force frequency variations, both

observed during the experiments. The maximum amplitude ratio for each solution with its

corresponding frequency is tabulated in Table 5.2.

As expected, the frequencies correspond closely with the values used to calibrate

the added mass term. The numerical solution does not reach the synchronous condition

observed in the experiments and its peak is reduced as the orifice size approached the

centerwell diameter. However, the analytical solution does reach this condition in all

orifice configurations. Figure 5.4 examines the results in the vicinity of the synchronous


Table 5.2, Maximum amplitude ratio comparison for non-linear damping only.
numerical solution analytical solution
AR o (rad/sec) Amp Ratio o (rad/sec) Amp Ratio
0.52 3.742 0.97 3.750 1.00
0.64 3.723 0.95 3.730 1.00
0.74 3.703 0.92 3.715 1.00
1.0 3.672 0.88 3.668 1.00

That the analytical solution is essentially a linear solution and it does reach the

synchronous condition suggests the linear damping term of the equation of motion cannot

be neglected and has a significant effect near the synchronous frequency.

A new least squares fit of the amplitude ratio data near the synchronous peak was

performed and the non-linear damping term (CL) of equation 3.22 was evaluated. Figure

5.5 shows the values obtained by this fit plotted against the orifice area ratio. An

excellent inverse square relationship was found to exist as illustrated in the figure.

Equation 5.10 describes the resulting expression for the linear damping coefficient.

CL 0 .06) S( cm) Eq. 5.10

The linear damping coefficient has units of one over time in equation 3.22.

Therefore scaling of this term should involve the natural frequency of the model and

prototype. In addition, equation 5.10 shows a further dependency on the draft of the

model. Such a scaling is described by equation 5.11 (M subscripts indicate the model

value and P the prototype).

Non-Linear Damping Amp Ratio Plot for AR = 0.74
+ maxima-minima analytic -- numeric


g 0.8

o 0.6

g 0.4


0.90 0.92 0.94 0.96 0.98 1.00
C/O /Co

Fig. 5.4, Amplitude ratio vs. frequency plot near the synchronous frequency for orifice
area ratio 0.74.

Plot of Linear Damping Coefficient vs. AR


h 0.21 ---
S0 1 y=0.0596x-2.519
R2 0.9974

a 0.11

.1 0.06

0.5 0.6 0.7 0.8 0.9 1.0

Fig. 5.5, Linear damping coefficient vs. orifice area ratio

0.06 1 SOMo Eq. 5.11
C, = o(l +c m) Sp EOq 5.11

Using this relationship, the non-linear damping terms were reevaluated. The no

orifice case gave a value of 1.1 for the fLID coefficient and the C' coefficient had a value

of 0.4 for all orifice ratios and no longer exhibited a frequency dependence.

The non-linear damping term in equation 3.22 has units of one over length. This

can be seen in the formulation of equations 5.5, 5.6 and 5.8 where the non-dimensional

resistance coefficients are divided by the model draft. Scaling by the draft should be

sufficient for the C' and the entrance and exit loss term, however thefL/D term contains a

draft to diameter ratio. Thus a more complex scaling involving both the diameter and

draft is required. Such a scaling is described below (M subscripts indicate the model value

and P the prototype).

1 1 So DIp SoM DipSoM

Equation 5.12 summarizes the least squares fit result for the non-linear term,

including scaling factors.

C, = DI o + 0.75+C )2So (l+cm) Eq. 5.12

f = 1.1, C' =0.4

Recalculating the numerical and analytical solutions with the above coefficients

results in the plots found in Figures 5.6 through 5.9. Table 5.3 shows the maximum

amplitude ratio comparison for these solutions. Again the agreement between analytical

and numerical amplitude ratio solutions is excellent. The phase plots show a sensitivity to

non-linear damping effects which the numerical solution reproduces well in the lower

frequencies. Overall, agreement with experimental data is excellent and neglecting us

appears to be justified.

Table 5.3, Maximum amplitude ratio comparison for combined
linear and non-linear damping.
numerical solution analytical solution
AR 0 (rad/sec) Amp Ratio o (rad/sec) Amp Ratio
0.52 3.750 1.00 3.750 1.00
0.64 3.731 1.00 3.731 1.00
0.74 3.719 1.00 3.715 1.00
1.0 3.691 0.98 3.691 1.00

Sensitivity of the Centerwell Heave Response Dynamics to Secondary Effects

As noted, there is good agreement of both the solutions with the experimental

amplitude ratio data over most of the frequency range. This does not indicate a

contribution of the ps added mass term in equation 3.22, which results from the

acceleration imparted by the orifice lip on the centerwell. The dimensions of the orifice lip

would seem to justify this conclusion and lend validity to the original assumption that the

effect of the orifice is only to increase the damping coefficient and therefore the coupling

of the spar and centerwell through the relative velocity. Indeed, the amplitude ratio plots

show a significant widening of the response curves with respect to frequency as the orifice

area decreases (i.e. as the bottom is closed off). In the extreme case of a closed bottom,

the amplitude ratio would necessarily be 1.0 for all frequencies and this is the limit

suggested by such widening of the amplitude ratio plots.

Full Eqn Amp Ratio Plot for Area Ratio = 0.52
+ maxima-minima analytic numeric



* 0.6

E 0.4


0.0 0.5 1.0 1.5 2.0 2.5 3.0
0o /o o

Full Eqn Phase Plot for Area Ratio = 0.52

+ maxima-minima analytic numeric


+30 +

G. 0

+ +

-90 4

0.0 0.5 1.0 1.5 2.0 2.5 3.0

Fig. 5.6, Solution plots for an orifice area ratio of 0.52.

Full Eqn Amp Ratio Plot for Area Ratio = 0.64
+ maxima-minima analytic -- numeric


" 0.8

0.2 --
S 0.4-


0.0 0.5 1.0 1.5 2.0 2.5 3.0
o)/C o

Full Eqn Phase Plot for Area Ratio = 0.64
+ maxima-minima analytic -- numeric


90- +7




0.0 0.5 1.0 1.5 2.0 2.5 3.0
a /o o

Fig. 5.7, Solution plots for an orifice area ratio of 0.64.

Full Eqn Amp Ratio Plot for Area Ratio = 0.74
+ maxima-minima analytic -- numeric

" 0.8

+ \

.2 + +
S 0.4 ----- j \ ---------

0.0 0.5 1.0 1.5 2.0 2.5 3.0
o/(C o

Full Eqn Phase Plot for Area Ratio = 0.74
+ maxima-minima --- analytic -- numeric

90 ----

n 60-
'I- +
30- +


0.0 0.5 1.0 1.5 2.0 2.5 3.0
-120 -
0.0 0.5 1.0 1.5 2.0 2.5 3.0

(0/O o

Fig. 5.8, Solution plots for an orifice area ratio of 0.74.

Full Eqn Amplitude Ratio Plot, No Orifice
+ maxima-minima analytic -- numeric


| 0.6

E 0.4

+ ++ +
0.0 0.5 1.0 1.5 2.0 2.5 3.0

0 /1 o

Full Eqn Phase Plot, No Orifice
+ maxima-minima analytic -- numeric


' 60
+ +o
<30 +

u+ +
-90 .


0.0 0.5 1.0 1.5 2.0 2.5 3.0

o o o o o o

Fig. 5.9, Solution plots for the no orifice case.

However, an examination of the effect of a term involving the spar acceleration is

necessary for completeness. Figure 5.10 shows the results of such a sensitivity test

conducted using the analytical solution. Recall that Ju = c,2 (+cm). The effect of

increasing the value of cm2 from 0.01 to 0.1 on the amplitude ratio and the phase difference

is enormous. This term causes the centerwell amplitude to exceed that of the spar motion

and shifts the entire phase plot up approximately 30 degrees.

Little difference exists between the plots for Cm2 values of 0.01 and 0. This lack of

sensitivity indicates that the assumption of a negligible contribution by the spar

acceleration term appears to be justified. A further check was performed by reducing the

value of Cm2 by an additional order of magnitude and allowing it to vary linearly with

frequency (i.e. c,2 = 0.001w). The result is a negligible difference with the amplitude

ratio plot for the c,2 = 0 case, but an interesting development in the phase plot. Figure

5.11 shows the phase plot for this case. Note the tendency of the phase to return to zero

degrees as the frequency increases which is observed in the experimental phase data.

Given the difficulty experienced in extracting the phase data and the previously

noted sensitivity of the phase, the result illustrated in Figure 5.11 cannot be considered

conclusive. The experimental variation at higher frequencies may very well be due to

experimental error or contamination of the data by secondary effects.

Another effect warranting investigation is that of the second order term associated

with the E coefficient of equation 3.22. Both numerical and analytical calculations were

performed by setting E = 0. The results are illustrated in Figure 5.12. As expected, the

absence of this term had little overall effect. The small peak in amplitude ratio at one half

the synchronous frequency and the decrease in the phase at the same frequency due to the

2c solution of Appendix A disappeared, indicating it was caused by the second order

term. The continued presence of the one third synchronous frequency phase decrease

indicates that this perturbation is not due to this term.

Finally, a least squares fit of the data over the entire frequency domain was

performed to obtain a solution for the case of linear damping only. The numerical and

analytical results of this solution are shown in Figure 5.13. Again, the results are as

expected. The solution has little effect on the amplitude ratio except for the removal of

the peak at one third synchronous frequency resulting from the 30 component of the non-

linear damping term. The same is true for the effect on the phase plot. Thus, the

perturbation at one third synchronous frequency is shown to be due to the non-linear

damping term.

Summary of Centerwell Heave Motion

The motion of the fluid in the centerwell is a result of the coupling between the

centerwell and the spar. This coupling is primarily the result of the surface forces which

derive from the relative velocities of the centerwell fluid and the spar itself. There is little

influence from the relative acceleration of the centerwell and spar.

When the centerwell moves in sync with the spar, no relative motion exists

between the centerwell fluid and the spar and the centerwell exhibits the behavior of an

undamped, freely oscillating mass-spring system. Though no actual coupling exists in this

condition, the centerwell amplitude ratio curves are very steep. Thus, with the centerwell

initially in the synchronous motion state, a slight drift of its motion away from this

condition will result in a rapid increase in relative velocity which will quickly restore the

centerwell motion to synchronous motion with respect to the spar.

Also, the centerwell amplitude ratio curves widen with respect to frequency if an

orifice is placed at the bottom of the centerwell. The amplitude ratio is only significant

over a very narrow frequency band near the synchronous frequency when no orifice is

present. This band gradually expands as the size of the opening at the bottom of the

centerwell decreases. The equations which describe the non-linear or viscous damping

coefficient would suggest that as the orifice lip becomes larger, the turbulence created at

the bottom of the centerwell increases and vortex formation similar to that described by

Knott and Flower increases, effectively closing off the bottom of the centerwell. That the

linear damping or inviscid pressure force term also plays an important role was

demonstrated by the fact that the numerical solution, which most accurately simulated the

non-linear effects, did not reach the synchronous condition without a linear damping term.

Thus, a combination of linear pressure forces and non-linear viscous or vortex

forming forces resulting from the relative velocity of the spar and centerwell fluid can

drive the centerwell fluid into a state where these forces disappear and the centerwell

oscillates as an undamped system. This is the basis of the synchronous motion observed in

the experiments.

The effect of these forces on a spar with a centerwell system moving under the

influence of waves will be examined in the next chapter.

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