Marine Hydrokinetic Turbine

Power-Take-Off Design for

Optimal Performance and Low
Impact on Cost-of-Energy
Preprint
M. Beam, B. Kline, B. Elbing, W. Straka, and
A. Fontaine
Pennsylvania State University

M. Lawson, Y. Li, and R. Thresher

National Renewable Energy Laboratory

M. Previsic
Re Vision Consulting, LLC
To be presented at the 32nd International Conference on Ocean,
Offshore and Arctic Engineering (OMAE 2013)
Nantes, France
June 9−14, 2013

NREL is a national laboratory of the U.S. Department of Energy, Office of Energy

Efficiency & Renewable Energy, operated by the Alliance for Sustainable Energy, LLC.

Conference Paper
NREL/CP-5000-58092
February 2013

Contract No. DE-AC36-08GO28308

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 Marine Hydrokinetic Turbine Power-Take-Off Design for Optimal Performance and Low
Impact on Cost-of-Energy

Mike J. Beam Brian l. Kline Brian E. Elbing

Applied Research Lab / Penn Applied Research Lab / Penn Applied Research Lab / Penn
State University State University State University
PO Box 30, State College, PA, PO Box 30, State College, PA, PO Box 30, State College, PA,
16804-0030 16804-0030 16804-0030

William Straka Arnold A. Fontaine Michael Lawson

ARL / PSU ARL / PSU Nat. Renewable Energy Lab
State College, PA, 16804-0030 State College, PA, 16804-0030 1617 Cole Blvd
Golden, CO 80401-3393

Ye Li Robert Thresher Mirko Previsic

Nat. Renewable Energy Lab Nat. Renewable Energy Lab Re VisionConsulting, LLC
Golden, CO 80401-3393 Golden, CO 80401-3393 Sacramento, CA 95831
mirko@re-vision.net

ABSTRACT (rotor or wave energy absorber design) and (2) the PTO design
Marine hydrokinetic devices are becoming a popular including the coupling drive train (bearing, seals, gearbox, etc.)
method for generating marine renewable energy worldwide. and the electric energy transmission system. The PTO design
These devices generate electricity by converting the kinetic process should attempt to optimize PTO performance with
energy of moving water, wave motion or currents, into consideration of overall device operation and impact on CoE.
electrical energy through the use of a Power-Take-Off (PTO) This design process should take into consideration device
system. Most PTO systems incorporate a mechanical or performance (Cp, operating parameters and predicted steady
hydraulic drive train, power generator and electric and unsteady operating loads), the operating environment,
control/conditioning system to deliver the generated electric overall PTO system efficiency and a trade-off study on system
power to the grid at the required state. Like wind turbine reliability, service life, maintenance schedule, weight and cost.
applications, the PTO system must be designed for high In addition, the impact of unit scale-up from one device to
reliability, good efficiency, and long service life with many in an array installation can be integrated into the device
reasonable maintenance requirements, low cost and an CoE.
appropriate mechanical design for anticipated applied steady While the PTO design protocol and its components are
and unsteady loads. The ultimate goal of a PTO design is high similar to wind turbine PTO systems, the design process must
efficiency, low maintenance and cost with a low impact on the additionally consider the marine environment requiring a more
device Cost-of-Energy (CoE). demanding design protocol. The PTO design for a marine
turbine must consider increased system loads in water as
INTRODUCTION opposed to air, a harsher operating environment (corrosion,
The long term success of hydrokinetic Marine Renewable water proofing or seals, marine fouling) and increased device
devices will be strongly coupled to their ability to produce deployment/accessibility costs. The presented PTO design
power in a competitive, cost-effective manner. A device’s CoE protocol leverages design experience in marine vehicle system
is primarily associated with the device’s overall efficiency, design and applies this to a 0.55 MW, horizontal axis
dependent on device design, and its manufacturing, deployment hydrokinetic turbine device for tidal operation. The design
and maintenance costs. The total device efficiency includes process included a CoE assessment study identifying key
contributions from (1) the energy extraction device design aspects of the PTO design that impact CoE.

1
 NOMENCLATURE DESIGN METHODOLOGY
Cp Power coefficient =P/(0.5ρU3) The PTO design process that will be described is in
PTO Power take off system reference to a NREL 0.55 MW hydrokinetic turbine designed
CoE Cost of energy under the US Dept. of Energy (DOE) Marine Renewable
rpm revolutions per minute Energy Reference Model program. This reference model
MW mega Watt program focused on developing a device design with: (1)
kW kilo Watt reasonable and validated power extraction performance, (2) a
m meter PTO system comprised of commercially available off-the shelf
(COTS) components (if available) and (3) a design that can be
manufactured and maintained with standard protocols. The goal
of the design was to produce a device with reasonable
performance specifications using common design process/tools
with commercially available components that can be easily
reproduced. The performance of the design would be validated
through testing/demonstrations with the aim of assessing device
design tools in general. Finally, a CoE analysis is performed to
estimate the device overall CoE for a single unit
implementation and the impact of multi-unit production and
deployment on CoE.
The goal of the reference model project was to develop a
simple robust design. So incorporating “radical turbine blade
designs” would have created large uncertainty in the
performance and CoE assessment, which is beyond the scope of
the project. While the effort could have focused on an
optimized device design to achieve high efficiency, such a
design would probably incorporate radical turbine blade
designs and unique PTO component designs specifically for
operation in the marine environment. This design process
would require an extensive system R&D design effort that
would increase costs, raise the CoE and result in possibly poor
reproducibility. An optimum turbine blade design may
incorporate design features that could increase manufacturing
costs substantially. Likewise, an optimized PTO design may
have increased cost, weight and size that could adversely affect
CoE through increased construction, deployment or
maintenance costs, more than offsetting any long-term CoE
reductions through performance gains. These features would
not be characteristic of a reference model. One goal of the
study was to utilize the CoE assessment to identify those
aspects/components of the device design with the greatest
impact on CoE.
The overall design process begins with close interactions
between the device design engineers and system design
engineers. The device specifications including power capacity,
size and deployment location are first defined. A hydrodynamic
design analysis is performed to define the rotor design (number
of blades, diameter, etc.) and operating specifications (rpm,
thrust, weight, anticipated steady and unsteady loads, etc.). For
this study, NREL developed a two turbine system providing a
combined 1.1 MW of power generation, see Lawson et al 1. The
turbine is designed to operate in the tidal resource of the
Admiralty Inlet in Puget Sound, WA, USA. The dual turbine
design is supported by a single tower anchored on the bed floor.
Figure 1. Schematic of the NREL reference turbine. A schematic of the overall system design is provided in Figure
1 with a summary of the pertinent physical and operational
parameters. In addition, unsteady rotor loads due to inflow

2
variations (turbulence, secondary flow components or time term fatigue failure) in commercial wind turbine applications. A
varying mass flow through the unit) 2- 4 or catastrophic events properly designed PTO system will take these unsteady loads
involving uneven shut down of one or both turbines was into account and attempt to limit vibration propagation through
considered in this analysis. These unsteady loads can be large the system.
with frequency content that must be understood to avoid The emphasis of the mechanical design of the PTO was to
vibration excitation near the device or system natural develop a robust design using commercially available
frequency. Flow induced vibration on the rotor plane must be components. This was done in an effort to preserve low overall
estimated as this will propagate through the drive train system cost and retain system maintainability. All drive train
impacting PTO system and component performance (efficiency, components would be housed within a water tight nacelle to
maintenance and life cycle). Unsteady loading is a primary maintain overall system buoyancy and provide a closed
driver in PTO system failure (short term catastrophic or long seawater tight system to permit use of standard industrial or
wind turbine components. The water tight nacelle design
eliminates the need to seal each component individually in a
water tight, seaworthy environment, provides easy access to the
tower and cross-arms eliminating the need for water proof
electrical connections and provides increased ballast volume for
buoyancy control. Figure 2 is a design image of the proposed
drive train assembly. The assembly consists of four major
systems: the bearing and seal assembly, the gearbox and
coupling section, generator section and the nacelle body. Each
system will be discussed in more detail later in this manuscript.
The power electronics for the two turbines would be located in
the system tower.

Drive Train Design

Generator
with Gearbox
The drive train design starts with the drive shaft
Cooling System & Shaft, specification. The current turbine design incorporates a variable
Couplings Bearings,
& Seals blade pitch system for controlling turbine rpm. A hollow drive
shaft design was sought to allow passage of power connectors
Rotor (input)
for the electric blade pitch actuators. It is desirable to minimize
drive shaft size while maintaining a performance specification
with a factor of safety of at least 1.5 to reduce overall weight,
and improve selection of supporting components such as seals,
bearings and couplings.
The drive shaft specified for this design has a shaft size
though the bearing bores of 317.5 mm with a 0.12 m bore
through the center. The shaft material selected is 17-4 ph
stainless steel. The material is specified to be at H1150 due to
its greatly enhanced resistance to stress corrosion cracking
when submersed in sea water. The factor of safety for the
material in condition H1150 is 8.2 during normal operation and
11. Seal and bearing assembly 11
2.5 taking into account the effect of the gusting moment. This
12. Rotor support weldment 12 design assumes that standard bearing locking nuts and hub
13. Fwd bulkhead
14. Seal and bearing flange
attachment methods will be used.
15. Center structure 13
The bearing and seal package was designed with the intent
16. Sliding drivetrain assembly
of protecting the balance of the drive train from sea water
14

15 elements, supporting the turbine rotor and withstanding subsea

gusting and unsteady loading which can propagate into the
16 drive train from the rotor plane. The operational parameters,
which the bearing and seal assembly would be subjected to, are
tabulated in Figure 1. The drive train design needs to support
the operational torque of the rotor and the system weight, the
weight of the rotor plane and the weights of the directly
connected couplings and components.
A modular bearing and seal package assembly was
Figure 2. PTO drive assembly. designed using commercially available standard components.

3
Emphasis was placed on selecting components which would expected to be approximately 7 watts. The energy consumption
minimize maintenance and maximize bearing package of the grease seals is expected to be of the same order of
component life. A modular bearing and seal package was magnitude.
proposed so that it could be assembled and tested separately Gearbox and Couplings Design Overview
from the balance of the assembly. Stock bearing and seal The gearbox was specified as a commercial design from
packages could be assembled and changed out as required thus manufacturers of gearboxes for wind turbines and hydro-
minimizing turbine operational down time. The use of turbines. The reference hydro-turbine had a rotor rpm of 10 to
commercially available components, minimizing maintenance 12. The low rpm and high torque must be carefully considered
and maximizing component availability would reduce this in a gearbox selection. The gearbox is used to transmit torque
contribution to the system CoE. from the rotor to a generator and is directly coupled to the drive
The resulting design concept, illustrated in Figure 3, train. Most commercial generators operate most efficiently at
incorporates the use of two Timken H961649/H961610 (317.5 rotational speeds of hundreds of rpm. As a result, the gearbox,
mm bore) tapered roller bearings. The Timken bearings were in this application, drives the generator as a speed increaser
selected due to their ability to sustain both radial and thrust rather than a speed reducer. The gearbox contains sets of
loads. With the use of two of these bearings an additional thrust planetary gears and parallel shaft spur gears arranged in groups
bearing is not required in the assembly. The bearings are grease to develop the required gear ratio. A typical gearbox contains
lubricated and with the loads specified have a predicted L10 three sets of reduction gears. These gear sets can be arranged to
life of 16.5 years. Horsepower consumption for both bearings at have a common center for the input shaft and output shaft, or
the designated shaft speed is nominally 15 Watts. The design they can be arranged to have a horizontal offset of the output
distance between the bearings is 1.83 m. The distance between shaft from a centered input shaft. One advantage of the offset
the centerline of the most outboard bearing and the center of shafts is to have a pass-through for electrical cables or
gravity of the rotor assembly is 1.52 m. hydraulic lines used for blade pitch control from the generator
The main sea water seal in the assembly is a mechanical side to the rotor side of the gearbox. This pass-through is
face seal (HSP-1014478-1, John Crane Inc.). The nominal shaft located on the centerline of the drive train, in-line with the rotor
size in the seal sleeve mounting is 311 mm. This seal, though shaft.
more expensive that comparably sized lip seals, was selected The gearbox transmits the torque through two connections
due to its outstanding projected life of 20 years. This projected to the drive train. On the low-speed turbine shaft side, a very
life cycle requires a two year maintenance interval for grease robust coupling is used to transmit the full torque of the drive
injection on the inboard side of the silicon carbide face seal to train, with some factor of safety for overload protection. On the
remove the minimal water seepage typical of this type of seal. high-speed generator side, a vibration dampening coupling is
The inboard grease seals (which do not touch sea water and are used. This coupling can also be designed with a torque-limiting
not subject to depth pressure) may be commercial lip seals. feature to protect the generator. A mechanical disc brake with
Energy consumption of the mechanical seal, due to friction, is hydraulic actuators can be located on the high-speed gearbox
side to provide controlled braking and to protect the drive
transmission components.
Gearbox Specifications and Designs: Extended Service
Life
The gearboxes that were considered for this study had
specific component design features. The planetary gears were
mounted on shafts with tapered roller bearings in an
arrangement called an “Integrated Flex-pin Bearing (IFB)” 5.
This technology was developed to minimize the angular
misalignment in the planet gears during transient torsional
overloads from the rotor. It also improved the life of the
planetary gears and permitted the gears and mounting
arrangements to be made smaller and lighter. , The gearbox
design included an auxiliary output shaft to drive one or more
hydraulic pumps. The hydraulic pumps were used for oil
lubrication for the gearbox. The auxiliary hydraulic power was
also used for blade pitch adjustments and for operating the
calipers for the disc brake. The gearbox lubrication systems
were chosen for longevity and needed for remote monitoring.
The systems included redundant oil filters and particulate
sensors to switch over to a new filter automatically as needed.
Multiple sensors were built into the gearboxes to sense
Figure 3 Bearing/seal package assembly. vibration levels, temperature and oil supply pressure for proper

4
lubrication. Oil cooling and conditioning were designed into the pointed out by Melfi et al. 8, the PM generators are appealing
lubrication system, with redundant components. Gearbox because they are efficient, reliable, and have improved
components, model and manufacturer, are provided in the performance (i.e. high power density, low power factor, low
maintenance and life cycle section of the paper. rotor temperature) and flexibility (i.e. synchronous operation).
The gearbox was specified with a closed-loop cooling The increased flexibility makes the operation frequency a
system that consisted of circulating lubrication oil through a degree of freedom in the system, which allows for operation at
heat exchanger 6. The heat exchanger consisted of a shell and base frequencies other than 50 or 60 Hz. However, there is a
tube design that was located in a pumping system tied into the price associated with this freedom, since the output frequency
generator cooling system, but in a separate loop. The heat must be 50 or 60 Hz an inverter must be added to the system. It
exchanger consisted of a series of cylindrical coils inside the was decided that the advantages of the PM generator
nacelle wall in a separate system from the generator cooling outweighed the increase in complexity of the electric
coils. Gearbox lubricating oil is pumped through the shell and conditioning system. The model selected was the ABB model
tube heat exchanger for cooling. The gearbox cooling system AMZ500LE10. Similarly, the remaining power conversion
consisted of a closed-loop pumping system with redundant components (transformer, drive control, cables, and connectors)
pumps and controls. The controls consisted of temperature were primarily selected based on available information and
regulators and sensors to maintain the gearbox at the correct were treated as off-the-shelf estimates for this conceptual
operating temperature. The gearbox cooling system was totally design. Generator and electric conditioning components, model
enclosed inside of the nacelle and was not subject to corrosion and manufacturer, are provided in the maintenance and life
from the surrounding seawater. For a cold startup, a gearbox cycle section of the paper.
lubrication oil heating system was used that bypassed the The synchronous induction generator 9 was designed with
cooling system with a set of control valves and temperature an integral cooling jacket. The generator design included a
sensors. cylindrical outer housing that contained a series of serpentine
Coupling Specifications and Designs: Extended Service Life channels where cooling fluid was pumped through the cooling
The couplings that were considered for this study included jacket. A heat exchanger in the form of a series of cylindrical
some design features that were required for long service life coils was located inside of the aft nacelle support housing along
and overload protection. The main rotor shaft is hollow to the inner surface of the steel shell. Cooling fluid was pumped
minimize weight and to provide a pass-through of electrical from the cooling jacket of the generator to the cylindrical coils
cables or hydraulic lines. The input coupling to the gearbox to facilitate the heat transfer to the surrounding seawater
would be an integrated style “shrink disc” coupling or a flexible through the steel wall of the aft nacelle support housing. The
coupling, depending on the ability to align the rotor shaft and generator cooling system consisted of a closed-loop pumping
the gearbox input shaft. The shrink disc coupling fits onto the system with redundant pumps and controls. The controls
input shaft of the gearbox requiring precise alignment during consisted of temperature regulators and sensors to maintain the
assembly of the rotor shaft. The assembly also requires special generator at the correct operating temperature. The generator
tools to properly torque the studs on the coupling to generate cooling system was totally enclosed inside of the nacelle and
the clamping forces required. The flexible couplings that were was not subject to corrosion from the surrounding seawater.
considered in the design incorporate a set of crowned gear
splines for both angular as well as offset misalignment. The
gear lubricant and seals were specified for long service life,
especially with the low speed of the rotor shaft and the marine
operating environment. The resilient couplings that were
considered for the generator or high-speed side of the gearbox
drive were designed for vibration control to protect the
generator. An additional feature was a “Multislip Torque
Limiter” 7 (2) that would provide torque limit to protect the
generator drive. The mechanical vibration control components
also act as electrical insulators to protect the drive from
electrical leakage current and electrical corrosion. An additional
feature of the resilient coupling components is the reduction of
the transmission of structure-borne noise in the drive.
Generator
The power conversion system greatly impacts the entire
design of the power train delivery system, and the choice of
using either an induction generator or a permanent magnet
(PM) generator could lead to two very distinct designs. While
induction generators are widely used and well understood, for Figure 4. Center support housing
the current MHK device a PM generator was selected. As

5
 Nacelle and Support Structure The aft conical support housing was designed to enclose
The support structure for the turbine was designed to house the generator in a water-tight enclosure and to provide access
the drive components and to act as the means of connection to on the aft end, shown in Figure 3. It consisted of a forward
the support arms on either side of the main support pile. The cylindrical flange that mated with the aft end of the center
center support housing of each nacelle acts as this support support housing and an aft flange that mated with the access
structure and is also designed to provide access to the gearbox, port on the aft end. The outer skin consisted of rolled steel
shown in Figure 4. The center support housing was designed plates with rolled “T-section” rings welded to the inside to
with cylindrical end flanges to mate with the forward and aft provide a water-tight pressure hull. The aft access port
nacelle conical support sections. Rectangular side flanges were consisted of a hemispherical shell with rolled “T-section” rings
designed to mate with the rectangular support arms. The center welded to the inside for support. The design combined strength
support housing could be made from cast steel or fabricated and rigidity with low weight. The design was in the preliminary
from steel as a weldment. The outer contour of the center stages and will require further refinement with FEA analysis to
support housing was made from thin rolled steel plate and optimize the structural components. The assembled support
could be attached to the center housing in split sections. Access structures are shown in Figures 7 and 8.
ports were located opposite the rectangular side flanges and at The nacelle was designed for easy system assembly and
other locations to provide access for inspection and access for maintenance. Access ports are provided in critical
maintenance. Bulkheads were designed to be located at the locations to enable necessary maintenance activities in an
forward and aft ends to mount the gearbox and to provide efficient manner. System build up is performed in center
support for the drive components. housing out fashion starting with the center conical support
The forward conical support housing was designed to housing and working outward aft and forward. The generator/
support the rotor and drive shaft with the bearing and seal gearbox assembly is mounted to a sliding rail system permitting
housing, shown in Figure 5. It consisted of an aft cylindrical easy access to these components through removal of the aft
bulkhead flange that mated with the forward end of the center conical section. Figure 2 illustrates a blow-out of this design.
support housing and a forward bulkhead flange that mated with Maintenance and Life Cycle
the forward end of the bearing and seal housing near the rotor The life cycle and preventative maintenance for large PTO
hub. Steel box beams were designed to tie the forward and aft system components can be comprehensive and can affect the
steel bulkhead flanges together into an integrated welded overall CoE for a marine hydrokinetic turbine. To estimate an
support structure. The outer skin consisted of rolled steel plates accurate CoE for the device, the maintenance and projected
with rolled “T-section” rings welded to the inside to provide a replacement requirements for the main PTO components were
water-tight pressure hull. This unified structure transferred the estimated based on manufacturer specifications
rotor forces (bending, axial and torsional) from the forward /recommendations and design/operation experience with similar
conical support housing to the center support housing. The components used in the marine environment. The following are
design combined strength and rigidity with low weight. The a statement of the data collected for each component.
design will require further refinement with FEA analysis to Information for the nacelle body was not available though a
optimize the structural components. major concern in a marine water environment would be bio
fouling.

Figure 5. Forward conical support housing. Figure 6. Aft conical support housing.

6
1. Mechanical face sea water seal
John Crane Seal provided by Hoffman Kane Distributors
Pittsburgh PA. Part # HSP-1014478-1.
Preventative: replace O rings, springs & primary carbon
ring and lap carbide mating ring every 2 years,
Component life: 20 or more years if used in clean water, 10
years of used in brown water.
2. Shaft Isolation Coupling – Gearbox to Generator
Renold Hi-Tec (direct from factory), Renold RB-3.86
$4769 net each (FOB Westfield, New York), est. finished
weight is 97 lb. No lubrication required, rubber elements
Estimated life of rubber elements – indefinite
3. Drive Shaft Bearings (2) Figure 7. Assembled support structures.
The Timken Company Canton Ohio, Part # H961649-
H961610
Preventative Maintenance: annual inspection, lubrication to
be provided via automated grease or oil lube system.
Component Life: Using a Timken on line program
assuming that the bearings are operated continuously a life
of 16.5 years is predicted.
NOTE: The life result is based on a report from the
Timken Tech Center date January 24, 2011;
Preliminary based on overhung load of "DRY" rotor
weight; pending a more detailed design & system analysis .
4. In board dry oil /grease seal (1)
SKF: V-ring Seal provided by Hoffman Kane Distributors Figure 8. Section through Assembled Support Structures.
Pittsburgh PA. Part # TBD, Price $974.
Component Life: Approximate life 2 years based on user
experience. 6. Generator
5. Rotor Drive Shaft (1) Manufacturer: ABB, Model: AMZ500LE10.
Preventative Maintenance: annual inspection, possible Preventive maintenance: Slip ring unit check every year/other
resurfacing in wear areas. maintain 5-10 year interval.
Component Life: Designed for stresses to be under the Lifecycle: over 10 to 20 years if properly maintained.
endurance limit of the material during normal operation. 7. Transformer
Design factor of safety in gust conditions is 2.5. Minimum Manufacturer: ABB, Type: 500 kVA Liquid Filled Pad mounted
design factor of safety during normal operation is 8. These Transformer Green-R-Pad FS or EFS.
factors of safety are based on overhung load of "DRY" Preventive maintenance: gauge readings, tank leaks, control
rotor weight; pending a more detailed design & system wiring & circuits, liquid dielectric test, temperature scan,
analysis including effects of dynamic loading. It should be insulator cleanliness need to be check once a year.
noted the that the drive train has been designed such that Lifecycle: with proper maintenance typically greater than 20
the shaft and bearings assume all radial loads generated at years.
the rotor plane in addition to torque and thrust. The 8. Drive Control
gearbox is subject to no extraneous rotor loads other than Manufacturer: ABB, Model: ACS800-17-0790-7.
torque. The housing containing this shaft, the bearings and Preventive maintenance: Heat sink temperature check &
the associated seals is in one self-contained assembly. This cleaning = 6 to 12 months/change cooling fan every 6
assembly can be removed as a completed assembly for years/capacitor change every 10 years.
ease of replacement and efficient maintenance with Component life: Fan life ~50,000 operating hours/capacitor
minimal down time. A spare assembly could be maintained lifespan = 45,000 (~5 years) to 90,000 hours.
such that the entire bearing and seal package could be
removed, inspected and any necessary items replaced or Cost & Economic Assessment of the PTO
repaired with periodic maintenance cycles. With the The lifecycle-cost analysis for the drive train was
current preliminary bearing selection having an estimated performed by RE Vision Consulting. In order to understand the
life of 16.5 years, this bearing & seal package exchange lifecycle cost of the powertrain, the following areas need to be
could be done every 2 to 4 years. understood: (1) Manufacturing cost at commercial production
levels, (2) regular maintenance requirements (i.e. oil change),
and (3) failure rates. Single unit pricing, while a useful

7
indicator for pilot cost, is a poor indicator for commercial scale
production runs. In order to provide an understanding of the
commercial system cost, an analogous cost assessment was
carried out to estimate the lifecycle cost from these systems
using wind-turbine cost data.
An important difference between a tidal turbine and a wind
turbine is the ease of access to the powertrain, which drives the
intervention requirements. Because maintenance access at sea
is extremely expensive, reducing the failure and maintenance
cycles is a key consideration in the design of these systems.
Redundancies were introduced in the design to reduce the
intervention interval to a minimum (1.1 annual repairs for a 2
rotor water turbine vs. 3.7 for a single rotor wind turbine with a
comparable power rating). While this measure increased the
total capital cost of the system by almost 20%, it reduces
overall lifecycle cost of the system by reducing the number of
times the device has to be recovered over its lifetime.
To characterize the cost of these intervention procedures, a
conceptual design for a dedicated vessel was established that
would allow the deployment and recovery of the powertrain
assembly. After reviewing handling requirements and the
potential of meeting those with vessels of opportunity, it was
determined that the most cost-effective approach would be to
have a custom designed vessel form an integral part of the plant
and staff it with a permanent crew. This is similar to the
approach taken by many leading tidal device manufacturers that
are realizing that the intervention requirement frequency and
the uniqueness of the environment in a tidal race would be best Figure 9: Capital cost of different drive-train configurations
addressed by use of specialty vessel that is equipped to meet the
requirements of the task. Total vessel cost was estimated at It is important to understand that the design of a powertrain
$13million to $16million. A crew of 20 men would be required cannot be carried out in isolation from the system design.
to safely operate the vessel in a tidal race. Other cost elements Lifecycle cost data from all systems were used to establish a
include (1) insurance, (2) replacement part cost, and (3) repair comprehensive lifecycle cost model. The resulting data was
personnel cost, (4) facilities and infrastructure. subsequently used to optimize the intervention intervals for the
With the exception of the sea water bearing and seal powertrain. While actual cost of electricity data for this project
assembly and the supporting structure for the PTO, the balance has not been released yet, the following illustration shows the
of the drive train uses existing wind turbine technology. To contribution of individual cost centers to the total system cost
estimate the cost of a commercially mature production level, (O&M costs are not shown in this figure). It shows that the
results from NREL’s WindPACT study were used and modified powertrain’s contribution to total cost will likely be significant
to take into account the following design differences between at commercial scale.
wind and water turbine designs:
Acknowledgments
1. Lower rotor shaft rpm resulting in increased gearbox cost The authors would like to acknowledge the support of the
2. Difference in rotor designs US Dept. of Energy and Michael Reed, Program Manager, for
3. Design redundancies in the drive train to minimize the support of this research design effort. DOE Research contracts
number of annual repairs for the water turbine #775396.
4. Sealing requirements
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