Design and Operational Performance of a
Standalone Passive Heave Compensation System for
a Work Class ROV
Andreas Huster, Hans Bergstrom, Jeff Gosior and Derek White
Oceanworks International Corp.
Burnaby, BC, Canada
{ahuster,hbergstrom,jgosior,dwhite}@oceanworks.com
Abstract—The design, testing and operational performance of
a passive flying-sheave type heave compensator are discussed.
The design process relied heavily on computer simulations of the
system dynamics. The paper presents a simple dynamic system
model that has sufficient detail to predict the performance of the
heave compensator and to guide the design process. The factory
testing and operational performance of the heave compensator,
as well as a possible extension to a hybrid active-over-passive
system, are described.
I. I NTRODUCTION
The number of Remotely Operated Vehicles (ROVs) in use
for offshore operations has surged as these underwater vehicles have become more and more indispensable for offshore
operations. With the high cost of offshore operations, ROV
operators are looking for greater utilization of their ROVs and
have asked for the ability to operate safely and efficiently at
higher and higher sea states. Increased vessel heave motion
at higher sea states makes operating an ROV progressively
more difficult and time-consuming. Further, at larger sea states,
vessel heave can cause slack lines and snap loading, which has
the potential to cause serious equipment damage and poses
a safety hazard to operators. These factors typically impose
an operational sea state limit for ROVs. When the sea state
exceeds this limit, the ROV cannot be operated and remains
idle.
To overcome this sea state limit, ROV launch and recovery
systems can be outfitted with heave compensation (HC) systems, which reduce the effect of vessel heave by shortening
and lengthening the effective length of the ROV umbilical in
concert with the vessel motion. There are several approaches to
heave compensation, including active winches, bobbing cranes
and flying sheave designs (see [1] for comparisons of various
heave compensation techniques). In this paper, we focus on
the flying sheave design, which has several advantages. Flying
sheave heave compensators can be packaged as compact,
standalone systems that require little or no power and are more
robust to failures.
These characteristics of the flying sheave design have permitted their application for retrofitting existing ROV launch
and recovery systems that have no heave compensation. Such
retrofits raise the effective sea state limit of existing launch
and recovery systems and extend their useful life. Even for
Fig. 1.
Model of the Passive Heave Compensator
new systems, the low-power, robust performance of the flying
sheave design provides an attractive alternative to active heave
compensation systems.
Oceanworks International Corp. (OWC) has recently designed, built, tested and commissioned a flying-sheave passive
heave compensation system. This paper describes the application and the design of the heave compensation system (see
Figure 1).
0-933957-38-1/09/$20.00 ©2009 MTS
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The purpose of this paper is to describe the engineering
process that enabled the development of this new passive heave
compensator. The design had to meet specific performance
criteria and was carried out on an accelerated schedule. This
required a deliberate design process and a good understanding
of the dynamic response of the heave compensator and the
overall ROV handling system. The design process was centred
on a dynamic simulation of the heave compensator as well
as the ROV, its tether management system, the umbilical
and vessel motions. This simulation was used to predict the
performance of the entire ROV handing system as a function
of various heave compensator design parameters, like friction,
inertia, accumulator size and hydraulic damping. This simulation capability was then used to perform design trade-offs,
such as the overall configuration of the heave compensator,
the allowable sheave friction and the accumulator stiffness.
Finally, the simulation was used to describe to the client the
expected operational performance of the heave compensator
at various operational conditions (i.e., vessel heave amplitude,
vessel heave period, and ROV depth).
Section II of the paper describes the application of the
heave compensator to retrofit an existing ROV launch and
recovery system. Section III describes the final design of
the heave compensation system and the rationale for specific
design features. Sections IV through VI provide details on the
dynamic simulation that was used to evaluate the design. The
paper describes a simulation approach that is sufficiently comprehensive to yield useful results, but not overly theoretical and
complicated. The dynamic simulation is based on earlier work
described in [2], [3].
Once the passive heave compensation system was fabricated, the analysis shifted from computer-based simulation
to hardware testing, to validate the earlier simulations and
to demonstrate adequate heave compensation performance.
The paper describes the test setup used for hardware testing
on the factory floor. For this test, the heave compensator
was instrumented with a variety of sensors to measure the
simulated vessel heave motion, the resulting umbilical tension
and the amount of heave compensation. Results from these
tests are presented in Section VII.
The heave compensator was recently commissioned into a
launch and recovery system for an ROV working in the Gulf of
Mexico. Section VIII describes early reports of its operational
performance. Section IX describes a hybrid active-over-passive
capability that could be used to improve the performance of the
heave compensator while retaining its robustness and operating
with substantially less power than a purely active system.
II. R ETROFIT A PPLICATION
While much of the discussion in this paper is generic,
the system that is described was designed to meet specific
requirements for a particular retrofit application. The existing
ROV system was a typical setup consisting of an ROV with
a tether management system (TMS), an A-frame and an
umbilical winch. The TMS is attached to the subsea end of
the umbilical and works as a submersed clump weight for the
Passive
Heave
Compensator
Umbilical
Winch
A-Frame
Vessel
TMS
ROV
Fig. 2. Passive Heave Compensator inserted between the A-Frame and the
Umbilical Winch of an Existing ROV LARS
umbilical. The ROV is connected to the TMS through a tether
that is spooled on the TMS. During launch and recovery, the
ROV is latched into the TMS. The A-frame is a crane that is
used to deploy the ROV over the side or stern of the surface
ship. The umbilical winch is used to spool several thousand
meters of umbilical. Together, the A-frame and the umbilical
winch constitute the Launch and Recovery System (LARS)
for the ROV System.
While a LARS without heave compensation is typical, it
has operational limitations during rough seas. When the sea
state exceeds the capability of the LARS, the ROV cannot
be operated. The passive heave compensator is intended to
augment the LARS by adding heave compensation without
having to replace any of the existing equipment. Figure 2
shows how a heave compensator can be added between the
A-frame and the umbilical winch of an existing ROV LARS.
The performance of a heave compensation system is often
expressed as a percentage. Two interpretations are possible:
ηposition
=
ηtension
=
residual motion of TMS
motion of vessel
tension amplitude with HC
1−
tension amplitude without HC
1−
(1)
(2)
The objective of this design was to achieve ηtension = 70%.
III. M ECHANICAL S YSTEM
The passive heave compensator consists of five sheaves, two
hydraulic cylinders, two 57 L accumulators and twelve 45 L
gas bottles (see Figure 1).
The sheaves are arranged in two sets. Three sheaves turn
on a fixed shaft mounted at the top of the structure and
two sheaves turn on a floating shaft at the bottom of the
structure. The floating shaft is constrained by linear guides
and two hydraulic cylinders. The umbilical enters the top of
the structure, wraps twice around the top and bottom sheaves,
and then exits on top.
The hydraulic cylinders have 1.35 m of travel. With two
wraps of umbilical, this provides 5.4 m of umbilical take-up
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Tb
1
2
A-Frame
Tc
Tf
bumb
3
Ac
Te
Td
kumb
Tg
6
4
5
xhc vhc
Fhc
mhc bhc
Ts
xv vv
Vessel
xs vs
TMS
ms bs
W
Fig. 3.
Dynamic Simulation Model
and the ability to compensate for up to ±2.7 m of vessel heave.
Each sheave has a sealed cylindrical roller bearing for reduced
friction and longer life.
The hydraulic cylinders are attached to two cross-ported
accumulators with connections on both the hydraulic piping
and the air piping. These links ensure that the two cylinders
and the two accumulators can each be modelled as one
larger device. The pressure in the accumulators is preset to
balance out the nominal load of the payload suspended on
the umbilical. Due to the weight of the umbilical, this preset
depends on the depth of the payload. The maximum nominal
load that the system can compensate is 19, 500 kg.
The preset pressure is adjusted so that the motion of the
flying sheaves is centred along the range of motion of the
cylinders. If the motion is not centred, the preset pressure can
be increased or decreased to compensate for the weight of the
deployed payload.
The system operates with air on the gas side of the accumulators. Ten of the gas bottles are connected to the accumulators.
This larger volume of air is required to increase the compliance
of the heave compensator (i.e., provide a softer spring) and to
reduce the tension variation between maximum and minimum
umbilical take-up.
The two remaining gas bottles are used to store pressurized
air at up to 415 bar. These storage bottles are large enough
to pressurize the whole system at shallower depths (i.e., low
pressure) or to increase the preset pressure during operation.
A 415 bar compressor is connected to the storage bottles and
can be used to recharge them during or after use.
IV. DYNAMIC S YSTEM M ODEL
A computer simulation of the dynamic behaviour of the
ROV system was an important part of the design process
for the passive heave compensator. This type of simulation
quickly highlights the most sensitive parameters in the design
and provides an early indication of the expected performance
of the compensator. Of interest is the level of compensation
that can be achieved with a given heave compensator design.
Figure 3 shows the model that was used for the dynamic
simulation. It shows the A-frame in the outboard position
and used to lift the deployed TMS. From the A-frame, the
umbilical passes through the heave compensator. The A-frame
and heave compensator are rigidly attached to the vessel,
which heaves up and down in response to the waves. The
umbilical winch is not included in the model because it does
not participate in the dynamics of interest. The umbilical is
shown to be attached to a fixed point on the vessel instead of
the umbilical winch.
The figure shows only the TMS, which assumes that the
ROV has been unlatched from the TMS. The unlatched configuration is more challenging because the mass, added mass
and drag values for the TMS alone are less than for the ROV
and TMS combined. Mass and drag of the submersed object
lead to higher tension forces, which tend to eclipse the nonidealities (e.g., friction and inertia) of the heave compensator.
In the simulation, xs and vs are the position and velocity
of the TMS, or submersed object, and ms and bs are the
effective mass (real and added) and the effective quadratic
drag coefficient. W is the weight of the submersed object.
The weight and mass of the umbilical are lumped in with the
submersed object, so ms and W depend on depth.
The equations of motion for the submersed object are:
ẋs
v̇s
= vs
1
=
(Ts − W − bs vs |vs |)
ms
(3)
(4)
The umbilical is modelled as a simple mass–spring–damper
system with the mass lumped into the submersed object. kumb
is the spring constant and is determined by a compliance
parameter γ that is expressed in percent-stretch per unit of
tension. An assumed damping coefficient ζumb is used to
compute the damping parameter from the spring constant and
the umbilical mass.
1
(5)
kumb =
γL
bumb = 2ζumb kumb mumb
(6)
The umbilical tension is related to the umbilical stretch ΔX
and its rate of change ΔV as follows:
Ts = kumb ΔX + bumb ΔV + W
(7)
ΔX and ΔV are defined below.
The umbilical runs across six sheaves. Sheave 1 is the
overboarding sheave mounted to the A-frame. Sheaves 2, 4
and 6 are the stationary sheaves of the heave compensator.
In this model, sheaves 1, 2, 4 and 6 are fixed to the vessel.
Sheaves 3 and 5 are the moving sheaves and are attached
to the main cylinders. The position of the moving sheaves is
xhc and their velocity is vhc . The cylinder pushes the sheaves
down with a force Fhc and the tension on the four legs of
the umbilical pull the sheaves up. Note that the dashed lines
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connecting Sheaves 2, 4 and 6 and Sheaves 3 and 5 in Figure 3
represent the common shafts.
Sheaves 2 through 6 are identical with mass msh , moment
of inertia Jsh and radius Rsh . The sheaves experience a sheave
loss βsh , expressed as a percent reduction of tension from the
leading side to the trailing side of the umbilical.
√ βsh is defined
for 90◦ wrap angles and is multiplied by 2 for 180◦ wrap
angles.
The model uses separate tensions for each of the segments
of umbilical, as indicated. For example, the tensions on either
side of Sheave 2 are related by the following equation:
Tc
= Tb (1 + σβsh )
(8)
where σ = +1 for clockwise rotation and σ = −1 for counterclockwise rotation.
The A-frame may have a different sheave, but for the
purpose of this paper, all the sheaves are assumed to be
identical.
The effective mass mhc for the linear motion of the heave
compensator includes several components, including the mass
of Sheaves 3 and 5, the mass massy attributed to the other
moving components of the heave compensator assembly (e.g.,
cylinder pistons, rods, clevises and the shaft), the mass moil
of hydraulic fluid and the effective linear mass attributed to
the rotational motion of each of the sheaves.
Consider the equivalent linear mass for the rotation of
Sheave 5.
Jsh
meq,sh =
(9)
2
Rsh
The equivalent linear mass of each of the other sheaves is
similar, except for multiples that capture how many units of
rotational motion correspond to one unit of linear motion, that
is 4× for Sheaves 1 and 2, 3× for Sheave 3, 2× for Sheave
4, 1× for Sheave 5 and none for Sheave 6.
mhc
=
(4 + 4 + 3 + 2 + 1)meq,sh + 2msh
+massy + moil
(10)
Fhc depends on the properties of the cylinder and accumulator system. The total piston area is AC and the preset
accumulator air pressure is p̄A with an air volume of V̄A .
Following the derivation in [3], this results in:
p̄A AC
(11)
Fhc = bhc vhc +
C
1− A
x
V̄A hc
The equations of motion for the heave compensator are:
= vhc
(12)
1
v̇hc =
(T c + T d + T e + T f − Fhc )
(13)
mhc
Finally, the vessel position is xv and its velocity is vv .
The whole system is tied together by the expression for the
umbilical stretch, which captures that every unit of heave
compensator motion is equivalent to four units of vessel or
submersed object motion.
ẋhc
ΔX
= xv − xs − 4xhc
(14)
ΔV
= vv − vs − 4vhc
(15)
V. C OMPUTER S IMULATION
The model described in the previous section has been
incorporated into a MATLAB simulation, where the response
to the system as a result of arbitrary vessel motion can be
computed. The vessel motion is modelled as a sine-wave
profile with amplitude ±1.2 m and an 8 s period.
The mass of the TMS is 3, 324 kg, its added mass is
estimated at 2, 834 kg and its in-water weight is 2, 106 kg.
The umbilical length for the simulation was chosen to be
600 m. This is a shallow depth where the umbilical has less
inherent compliance. At this depth, the umbilical mass is
4.29 kg/m × 600 m = 2, 574 kg and its weight in water is
4.1 kg/m × 600 m = 2, 460 kg. Thus,
ms
W
=
=
8, 732 kg
4, 566 kg × 9.81 N/kg = 44, 700 N.
(16)
(17)
The 600 m depth was chosen so that this value of W matches
the nominal tension of the factory testing results presented in
Section VII.
The quadratic drag of the TMS is given by:
=
1
ρCd Atms = 2, 250 N/(m/s)2
2
ρ
=
1000 kg/m3
(19)
Cd
=
1.2
(20)
bs
where:
Atms
= πR2 = π(1.1m)2 = 3.8 m2
(18)
(21)
The umbilical compliance is γ = 1.2410−8 N −1 and the
assumed damping factor is ζumb = 0.06 . With mumb =
2, 574 kg,
kumb
bumb
=
=
1.34 × 105 N/m
3
2.23 × 10 N/(m/s).
(22)
(23)
The sheaves have the following parameters:
Rsh
=
0.782 m
(24)
msh
=
=
188 kg
56.7 kg · m2
(25)
(26)
=
0.6 %
(27)
Jsh
βsh
The size of the sheaves is governed by the minimum bend
radius of the umbilical and their strength is governed by the
tension rating of the system. The design uses roller bearings
specifically to keep the sheave losses as low as possible.
The value for βsh was determined during the factory testing
described in Section VII.
With massy = 946 kg and moil = 59.1 kg, the effective
accelerated mass of the heave compensator is:
mhc
=
2, 681 kg
(28)
The two cylinders, taken together, have a piston area of
AC = 0.0497 m2 . When the heave compensator is centred,
the air volume contained in the accumulator and the ten gas
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1.5
Tension, HC Off
Tension, HC On
Static Tension
4
5.4
1
x 10
5.2
5
4.8
Tension (N)
Position (m)
0.5
0
−0.5
−1.5
0
5
10
15
20
Time (s)
25
30
35
4.4
4.2
vessel
sub
HC * 4
Umb
−1
4.6
4
3.8
40
3.6
(a) Heave Compensator Disabled
0
5
10
15
20
Time (s)
25
30
35
40
1.5
Fig. 5.
Simulation Results for Tension
1
Position (m)
0.5
0
−0.5
vessel
sub
HC * 4
Umb
−1
−1.5
0
5
10
15
20
Time (s)
25
30
35
40
(b) Heave Compensator Enabled
Fig. 4.
Simulation Results for Position and Umbilical Stretch
bottles is V̄A = 0.490 m3 . The preset pressure depends on the
nominal umbilical tension:
4W
= 3.61 M P a
(29)
p̄A =
AC
Hydraulic damping in the heave compensator is based on an
assumed pressure drop of Bhc = 2 M P a/(m3 /s), which leads
to:
bhc
= Bhc A2C = 4.93 × 103 N/(m/s)
(30)
VI. S IMULATION R ESULTS
This section presents the results of two 40 s dynamic
simulations of the ROV system, with and without the heave
compensator enabled. Simulations such as these were a key
design tool for the passive heave compensation system. Figure 4 shows how the various elements of the ROV system
moved and Figure 5 shows the umbilical tension.
Figure 4 contains two plots. Plot (a) shows the performance
of the system when the heave compensator is disabled, that is,
before the installation of the heave compensation capability.
Plot (b) shows the performance after the heave compensator
has been installed. Both of these plots show the motion of
the vessel (the input to the simulation), the motion of the
TMS (submersed mass or sub), the motion of the heave
compensator, and the stretch of the umbilical.
Note that the vessel motion is the same on both plots.
Without the heave compensator, the TMS tracks the motion of
the vessel, as the system’s only compliance is in the umbilical.
When the heave compensator is added in the second plot,
the motion of the TMS is significantly reduced, as the heave
compensator is able to provide much more compliance. The
heave compensator motion increases from nothing in the first
plot, where the heave compensator is disabled, to reasonably
good tracking of the vessel motion. Note that the heave
compensator motion is shown with an amplification of 4, as
every unit of actual heave compensator motion corresponds
to 4 units of motion of the other parts. These plots also
show that when the heave compensator provides the necessary
compensation, the umbilical stretch, which is closely related
to umbilical tension, is significantly reduced.
Figure 5 shows the umbilical tensions for both cases on the
same plot. The reduction of variations in the umbilical tension
as a result of adding the heave compensator to the system can
be seen on this plot. Note that this reduction corresponds to
the reduction in motion of the TMS. This is not surprising
because the vessel-induced variations of the umbilical tension
actually cause this motion.
Ideally, the tension variations on the umbilical could be
eliminated completely. This could be achieved if the heave
compensator motion were able to track perfectly the motion
of the vessel. This is not possible for a passive heave compensation system as it has several inherent physical limitations that
prevent perfect tracking. Key limitations are sheave losses and
the inertia of the sheaves and other moving parts. As Figure 4b
shows, even when the heave compensator preforms reasonably
well at tracking the vessel motion, the motion of the TMS has
not been eliminated.
In this particular simulation, the reduction in TMS motion
is ηposition = 65% and the reduction in tension variations is
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Load Cell
T
Test
Mass
x1
x2
Passive
Heave
Compensator
Test
Cylinder
Fig. 7.
Fig. 6.
Factory Testing
ηtension = 68%. These results are typical for passive heave
compensation systems and are near the design objective.
Simulations like this were used throughout the design
phase of the heave compensator to evaluate different design
choices. A key finding was the importance of roller bearings
instead of journal bearings for the heave compensator sheaves.
Journal bearings have substantially higher sheave losses and
could result in a heave compensator with little or no benefit.
Reducing inertia was also found to be important, which led
to the optimized reduction of the sheave mass. Finally, the
air volume has to be chosen to ensure that the heave compensator compliance is sufficiently greater than the expected
compliance of the umbilical. This evaluation was performed
with the computer simulation.
VII. FACTORY T ESTING
This section describes the procedure and equipment used for
dynamic testing of the heave compensator at the factory after
assembly. The objective of the test was to confirm the ability
of the passive heave compensator to minimize variations in
the umbilical tension induced by simulated vessel motion.
The dynamic test setup (see Figures 6 and 7) uses a
hydraulic cylinder to simulate the vessel motion. A test mass
was suspended on a wire rope to simulate the combined mass
of the TMS and the umbilical. The wire rope was connected
through the heave compensation system to a hydraulic test
cylinder with a 2:1 pulley arrangement. The usable stroke of
the test cylinder was 1.2 m to permit ±1.2 m of wire rope
travel. The flow rate of the test cylinder was controlled by an
operator with a manual forward-off-reverse valve to simulate
Test Setup
the vessel motion. A test mass of 4, 540 kg was selected to
correspond to the nominal tension of the computer simulations
discussed in the previous section.
Testing was performed in two different modes: with the
heave compensator disabled and with it enabled. For all tests,
the wire rope tension (T in Figure 7) and the displacements of
the test cylinder (x1 ) and the test mass (x2 ) were recorded. The
results from both experiments were compared to determine the
effectiveness of the heave compensator.
Although the test setup and the computer simulations are
matched in terms of umbilical tension, the payloads have
different effective masses. The simulations are based on ms =
8, 732 kg (see Section V), which is almost twice the test mass,
as it incorporates added mass and the reduced weight in water.
Also, the test setup has no hydrodynamic drag.
This test setup differs from an operational deployment in
three significant aspects. First, the ratio of the effects that drive
the heave compensator (e.g., effective mass of the payload and
drag on the payload) over the effects that inhibit the heave
compensator (e.g., sheave losses due to umbilical tension) is
smaller for the test setup. Thus, performance of the heave compensator during an actual deployment is expected to be better
than during the factory tests. Second, the test setup uses a short
piece of umbilical with proportionally less compliance than an
at-sea deployment. Third, the simulation of the vessel motion
was limited by the available hydraulic controls, which was a
simple forward-off-reverse valve controlled by an operator.
Figures 8 and 9 show some results from the factory testing
in a format similar to the simulation results. The vessel motion
is now the extension of the hydraulic cylinder (x1 ), modified
for the 2:1 pulley arrangement. The sub motion is the test mass
(x2 ). The HC * 4 motion is the difference between vessel and
sub when the heave compensator is enabled, zero otherwise.
The tension plots show the measurements from the load cell.
Figure 8 shows that the simulated vessel motion is created
with a series of constant-speed segments, a limitation imposed
by the simple forward-off-reverse control of the hydraulic
cylinder. The tension plots are significantly affected as all
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vessel
sub
HC * 4
1.4
1.2
1
Position (m)
0.8
0.6
0.4
0.2
0
−0.2
−0.4
−0.6
0
5
10
15
20
Time (s)
25
30
35
40
(a) Heave Compensator Disabled
vessel
sub
HC * 4
1
0.8
0.6
the acceleration, and with it the tension variations, is focused
onto the transitions in the motion profile. The resulting highfrequency content causes the observed ringing in the tension
plots.
However, Figure 9 clearly shows the reduction in the size
of the tension variations that can be attributed to enabling
the heave compensator. The RMS tension variation is reduced
from 9, 700 N to 3, 000 N , resulting in ηtension = 68%, which
is near the design objective.
Position (m)
0.4
0.2
0
−0.2
−0.4
−0.6
−0.8
0
5
10
15
20
Time (s)
25
30
35
40
(b) Heave Compensator Enabled
Fig. 8.
Factory Testing Results for Position and Umbilical Stretch
4
9
x 10
Tension, HC Off
Tension, HC On
Static Tension
8
7
Tension (N)
6
5
VIII. O PERATIONAL P ERFORMANCE
The passive heave compensation system has now been
deployed at sea in the Gulf of Mexico for several months.
It was installed as part of a retrofit for an ROV Launch
and Recovery System, adding heave compensation to the
system. Figure 10 shows the heave compensator installed
on the ROV support vessel. Early reports indicate that the
heave compensator is meeting its design objectives. Heave
compensation has extended the operational capability of the
ROV and has allowed the ROV to be operated through the
winter months. With the passive heave compensator, the ROV
can be operated in 3.5 m seas, which was not possible before.
In 2 m seas, the heave compensator has reduced the TMS
motion to below 0.2 m and eliminated snap loading.
IX. ACTIVE - OVER -PASSIVE E XTENSION
4
3
2
1
Fig. 10. The Passive Heave Compensator installed on a vessel as part of an
ROV Launch and Recovery System
0
5
10
Fig. 9.
15
20
Time (s)
25
30
35
Factory Testing Results for Tension
40
This passive heave compensator design is readily extensible
to a hybrid active-over-passive system, which preserves the
robustness and standalone features of the passive design, but
improves the compensation performance and does so at a fraction of the power required for an active heave compensating
umbilical winch. Such a system has already been proposed
by [4].
To achieve active-over-passive performance, the heave compensator requires an actuator that can inject forces in parallel
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with the existing hydraulic cylinder. This actuator is used to
overcome the main limitations of the passive heave compensator, like sheave friction, inertia and phase mismatch. The key
advantage of this approach is that the required force and power
capacity of this actuator is substantially less than that required
for a purely active system, because the actuator is responsible
only for the difference between the passive performance (e.g.,
70%) and the desired active performance (e.g., 90%).
The actuator is controlled based on several sensor inputs.
Possible sensors include vessel motion, TMS motion, umbilical tension and heave compensator motion. The control algorithm can be tailored to the specific performance objectives
and the sensor measurements that are available on the system.
Active systems inherently have more failure modes than
passive systems, and this general rule also applies to the activeover-passive hybrid system. However, if a failure disables the
active capability, an active-over-passive system can retain its
passive capability. This would lead to a reduction in performance, but not a loss of the heave compensation capability,
which is a key operational advantage of an active-over-passive
design.
X. C ONCLUSION
This paper has described the design process, factory testing
and operational performance of a passive heave compensation
system. A key component of the design process for the passive heave compensator were dynamic simulations of various
candidate designs. The paper describes a computer simulation
of the actual design and highlights some of the key findings.
Although the dynamic model for these computer simulations
is quite simple, this type of simulation is able to direct the
design effort to those aspects of the design that are most likely
to increase the overall performance.
These simulations indicated that the heave compensator
design should utilize the highest efficiency bearings, in order
to reduce the tension losses through the heave compensator,
and have low inertia. They also established expectations of
operational performance for the passive heave compensator.
The paper described factory testing of the heave compensator. Because at-sea conditions are difficult to replicate in the
factory, a simplified test setup was used to check the most
important aspects of the heave compensator performance and
to verify the simulation model. After factory testing, the heave
compensator was installed on an ROV support vessel, where
early reports indicate that it is meeting its design objectives.
Performance of the heave compensator can be further improved by a possible extension to create a hybrid active-overpassive system. The advantages of this type of system were
discussed.
R EFERENCES
[1] J. E. Adamson, “Efficient heave motion compensation for
cable-suspended systems.” Underwater Intervention, 2003. Available online: http://www.oceanworks.com/cms/pdfs/OW2003 Heave%
20Compensation.pdf
[2] F. R. Driscoll, M. Nahon, and R. G. Lueck, “A comparison of shipmounted and cage-mounted passive heave compensation systems,” Journal of Offshore Mechanics and Arctic Engineering, vol. 122, no. 3, pp.
214–221, 2000. DOI: 10.1115/1.1287167
[3] A. Huster, A. Dayani, and D. Lo, “Design and testing of a snap load
alleviator for a submarine rescue vehicle handling system,” in Proceedings
of the Oceans 2007 Conference. Vancouver, Canada: MTS/IEEE, 2007.
DOI: 10.1109/OCEANS.2007.4449301
[4] J. D. Haney, D. W. Carey, and Q. Z. Rhodes, II, “Low power/high
performance active heave compensation,” Sea Technology Magazine,
July 2002. Available online: http://findarticles.com/p/articles/mi qa5367/
is 200207/ai n21315906
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