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.

Contract No. DE-AC36-08GO28308

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

Conference Paper

NREL/CP-5000-58092

February 2013
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1

Marine Hydrokinetic Turbine Power-Take-Off Design for Optimal Performance and Low 

Impact on Cost-of-Energy

Mike J. Beam

Applied Research Lab / Penn 

State University

PO Box 30, State College, PA, 

16804-0030

Brian l. Kline

Applied Research Lab / Penn 

State University

PO Box 30, State College, PA, 

16804-0030

Brian E. Elbing

Applied Research Lab / Penn 

State University

PO Box 30, State College, PA, 

16804-0030

William Straka

ARL / PSU

State College, PA, 16804-0030

Arnold A. Fontaine

ARL / PSU

State College, PA, 16804-0030

Michael Lawson

Nat. Renewable Energy Lab

1617 Cole Blvd

Golden, CO 80401-3393

Ye Li 

Nat. Renewable Energy Lab

Golden, CO 80401-3393

Robert Thresher

Nat. Renewable Energy Lab

Golden, CO 80401-3393

Mirko Previsic

Re VisionConsulting, LLC

Sacramento, CA 95831

mirko@re-vision.net

ABSTRACT

Marine hydrokinetic devices are becoming a popular 

method for generating marine renewable energy worldwide.

These devices generate electricity by converting the kinetic 

energy of moving water, wave motion or currents, into 

electrical energy through the use of a Power-Take-Off (PTO) 

system. Most PTO systems incorporate a mechanical or 

hydraulic drive train, power generator and electric 

control/conditioning system to deliver the generated electric 

power to the grid at the required state. Like wind turbine 

applications, the PTO system must be designed for high 

reliability, good efficiency, and long service life with 

reasonable maintenance requirements, low cost and an 

appropriate mechanical design for anticipated applied steady 

and unsteady loads. The ultimate goal of a PTO design is high 

efficiency, low maintenance and cost with a low impact on the 

device Cost-of-Energy (CoE).

INTRODUCTION

The long term success of hydrokinetic Marine Renewable 

devices will be strongly coupled to their ability to produce 

power in a competitive, cost-effective manner. A device’s CoE 

is primarily associated with the device’s overall efficiency, 

dependent on device design, and its manufacturing, deployment 

and maintenance costs. The total device efficiency includes 

contributions from (1) the energy extraction device design 

(rotor or wave energy absorber design) and (2) the PTO design 

including the coupling drive train (bearing, seals, gearbox, etc.) 

and the electric energy transmission system. The PTO design 

process should attempt to optimize PTO performance with 

consideration of overall device operation and impact on CoE.

This design process should take into consideration device 

performance (Cp, operating parameters and predicted steady 

and unsteady operating loads), the operating environment, 

overall PTO system efficiency and a trade-off study on system 

reliability, service life, maintenance schedule, weight and cost.

In addition, the impact of unit scale-up from one device to 

many in an array installation can be integrated into the device 

CoE.

While the PTO design protocol and its components are 

similar to wind turbine PTO systems, the design process must 

additionally consider the marine environment requiring a more 

demanding design protocol. The PTO design for a marine 

turbine must consider increased system loads in water as 

opposed to air, a harsher operating environment (corrosion, 

water proofing or seals, marine fouling) and increased device 

deployment/accessibility costs. The presented PTO design 

protocol leverages design experience in marine vehicle system 

design and applies this to a 0.55 MW, horizontal axis 

hydrokinetic turbine device for tidal operation. The design 

process included a CoE assessment study identifying key 

aspects of the PTO design that impact CoE.
2

NOMENCLATURE

Cp Power coefficient =P/(0.5ρU3

)

PTO Power take off system

CoE Cost of energy

rpm revolutions per minute

MW mega Watt

kW kilo Watt

m meter

DESIGN METHODOLOGY

The PTO design process that will be described is in 

reference to a NREL 0.55 MW hydrokinetic turbine designed 

under the US Dept. of Energy (DOE) Marine Renewable 

Energy Reference Model program. This reference model 

program focused on developing a device design with: (1) 

reasonable and validated power extraction performance, (2) a 

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 al1

. 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.

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 

Figure 1. Schematic of the NREL reference turbine.
3

variations (turbulence, secondary flow components or time 

varying mass flow through the unit)2- 4 or catastrophic events 

involving uneven shut down of one or both turbines was 

considered in this analysis. These unsteady loads can be large 

with frequency content that must be understood to avoid 

vibration excitation near the device or system natural 

frequency. Flow induced vibration on the rotor plane must be 

estimated as this will propagate through the drive train 

impacting PTO system and component performance (efficiency, 

maintenance and life cycle). Unsteady loading is a primary 

driver in PTO system failure (short term catastrophic or long 

term fatigue failure) in commercial wind turbine applications. A 

properly designed PTO system will take these unsteady loads 

into account and attempt to limit vibration propagation through 

the system.

The emphasis of the mechanical design of the PTO was to 

develop a robust design using commercially available 

components. This was done in an effort to preserve low overall 

system cost and retain system maintainability. All drive train 

components would be housed within a water tight nacelle to 

maintain overall system buoyancy and provide a closed 

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

The drive train design starts with the drive shaft 

specification. The current turbine design incorporates a variable 

blade pitch system for controlling turbine rpm. A hollow drive 

shaft design was sought to allow passage of power connectors 

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 

2.5 taking into account the effect of the gusting moment. This 

design assumes that standard bearing locking nuts and hub 

attachment methods will be used.

The bearing and seal package was designed with the intent 

of protecting the balance of the drive train from sea water 

elements, supporting the turbine rotor and withstanding subsea 

gusting and unsteady loading which can propagate into the 

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. 

Rotor (input)

Generator

with 

Cooling System Gearbox

&

Couplings

Shaft,

Bearings,

& Seals

11

12

13

14

15

16

11. Seal and bearing assembly

12. Rotor support weldment

13. Fwd bulkhead

14. Seal and bearing flange

15. Center structure

16. Sliding drivetrain assembly
4

Emphasis was placed on selecting components which would 

minimize maintenance and maximize bearing package 

component life. A modular bearing and seal package was 

proposed so that it could be assembled and tested separately 

from the balance of the assembly. Stock bearing and seal 

packages could be assembled and changed out as required thus 

minimizing turbine operational down time. The use of 

commercially available components, minimizing maintenance 

and maximizing component availability would reduce this 

contribution to the system CoE.

The resulting design concept, illustrated in Figure 3, 

incorporates the use of two Timken H961649/H961610 (317.5 

mm bore) tapered roller bearings. The Timken bearings were 

selected due to their ability to sustain both radial and thrust 

loads. With the use of two of these bearings an additional thrust 

bearing is not required in the assembly. The bearings are grease 

lubricated and with the loads specified have a predicted L10 

life of 16.5 years. Horsepower consumption for both bearings at 

the designated shaft speed is nominally 15 Watts. The design 

distance between the bearings is 1.83 m. The distance between 

the centerline of the most outboard bearing and the center of 

gravity of the rotor assembly is 1.52 m.

The main sea water seal in the assembly is a mechanical 

face seal (HSP-1014478-1, John Crane Inc.). The nominal shaft 

size in the seal sleeve mounting is 311 mm. This seal, though 

more expensive that comparably sized lip seals, was selected 

due to its outstanding projected life of 20 years. This projected

life cycle requires a two year maintenance interval for grease 

injection on the inboard side of the silicon carbide face seal to 

remove the minimal water seepage typical of this type of seal. 

The inboard grease seals (which do not touch sea water and are 

not subject to depth pressure) may be commercial lip seals.

Energy consumption of the mechanical seal, due to friction, is 

expected to be approximately 7 watts. The energy consumption 

of the grease seals is expected to be of the same order of 

magnitude.

Gearbox and Couplings Design Overview

The gearbox was specified as a commercial design from

manufacturers of gearboxes for wind turbines and hydroturbines. The reference hydro-turbine had a rotor rpm of 10 to 

12. The low rpm and high torque must be carefully considered 

in a gearbox selection. The gearbox is used to transmit torque 

from the rotor to a generator and is directly coupled to the drive 

train. Most commercial generators operate most efficiently at 

rotational speeds of hundreds of rpm. As a result, the gearbox, 

in this application, drives the generator as a speed increaser

rather than a speed reducer. The gearbox contains sets of 

planetary gears and parallel shaft spur gears arranged in groups 

to develop the required gear ratio. A typical gearbox contains 

three sets of reduction gears. These gear sets can be arranged to 

have a common center for the input shaft and output shaft, or 

they can be arranged to have a horizontal offset of the output 

shaft from a centered input shaft. One advantage of the offset 

shafts is to have a pass-through for electrical cables or 

hydraulic lines used for blade pitch control from the generator 

side to the rotor side of the gearbox. This pass-through is 

located on the centerline of the drive train, in-line with the rotor 

shaft.

The gearbox transmits the torque through two connections 

to the drive train. On the low-speed turbine shaft side, a very 

robust coupling is used to transmit the full torque of the drive 

train, with some factor of safety for overload protection. On the 

high-speed generator side, a vibration dampening coupling is 

used. This coupling can also be designed with a torque-limiting 

feature to protect the generator. A mechanical disc brake with 

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 

vibration levels, temperature and oil supply pressure for proper Figure 3 Bearing/seal package assembly.
5

lubrication. Oil cooling and conditioning were designed into the 

lubrication system, with redundant components. Gearbox 

components, model and manufacturer, are provided in the 

maintenance and life cycle section of the paper.

The gearbox was specified with a closed-loop cooling 

system that consisted of circulating lubrication oil through a 

heat exchanger6

. The heat exchanger consisted of a shell and 

tube design that was located in a pumping system tied into the 

generator cooling system, but in a separate loop. The heat 

exchanger consisted of a series of cylindrical coils inside the 

nacelle wall in a separate system from the generator cooling 

coils. Gearbox lubricating oil is pumped through the shell and 

tube heat exchanger for cooling. The gearbox cooling system 

consisted of a closed-loop pumping system with redundant 

pumps and controls. The controls consisted of temperature 

regulators and sensors to maintain the gearbox at the correct 

operating temperature. The gearbox cooling system was totally 

enclosed inside of the nacelle and was not subject to corrosion 

from the surrounding seawater. For a cold startup, a gearbox 

lubrication oil heating system was used that bypassed the 

cooling system with a set of control valves and temperature 

sensors.

Coupling Specifications and Designs: Extended Service Life

The couplings that were considered for this study included

some design features that were required for long service life 

and overload protection. The main rotor shaft is hollow to 

minimize weight and to provide a pass-through of electrical 

cables or hydraulic lines. The input coupling to the gearbox 

would be an integrated style “shrink disc” coupling or a flexible 

coupling, depending on the ability to align the rotor shaft and 

the gearbox input shaft. The shrink disc coupling fits onto the 

input shaft of the gearbox requiring precise alignment during 

assembly of the rotor shaft. The assembly also requires special 

tools to properly torque the studs on the coupling to generate 

the clamping forces required. The flexible couplings that were 

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 

the current MHK device a PM generator was selected. As 

pointed out by Melfi et al.

8

, the PM generators are appealing 

because they are efficient, reliable, and have improved 

performance (i.e. high power density, low power factor, low 

rotor temperature) and flexibility (i.e. synchronous operation). 

The increased flexibility makes the operation frequency a 

degree of freedom in the system, which allows for operation at 

base frequencies other than 50 or 60 Hz. However, there is a 

price associated with this freedom, since the output frequency 

must be 50 or 60 Hz an inverter must be added to the system. It 

was decided that the advantages of the PM generator 

outweighed the increase in complexity of the electric 

conditioning system. The model selected was the ABB model 

AMZ500LE10. Similarly, the remaining power conversion 

components (transformer, drive control, cables, and connectors) 

were primarily selected based on available information and 

were treated as off-the-shelf estimates for this conceptual 

design. Generator and electric conditioning components, model 

and manufacturer, are provided in the maintenance and life 

cycle section of the paper.

The synchronous induction generator9 was designed with 

an integral cooling jacket. The generator design included a 

cylindrical outer housing that contained a series of serpentine 

channels where cooling fluid was pumped through the cooling 

jacket. A heat exchanger in the form of a series of cylindrical 

coils was located inside of the aft nacelle support housing along 

the inner surface of the steel shell. Cooling fluid was pumped 

from the cooling jacket of the generator to the cylindrical coils 

to facilitate the heat transfer to the surrounding seawater 

through the steel wall of the aft nacelle support housing. The 

generator cooling system consisted of a closed-loop pumping 

system with redundant pumps and controls. The controls 

consisted of temperature regulators and sensors to maintain the 

generator at the correct operating temperature. The generator 

cooling system was totally enclosed inside of the nacelle and 

was not subject to corrosion from the surrounding seawater.

Figure 4. Center support housing
6

Nacelle and Support Structure

The support structure for the turbine was designed to house 

the drive components and to act as the means of connection to 

the support arms on either side of the main support pile. The 

center support housing of each nacelle acts as this support 

structure and is also designed to provide access to the gearbox, 

shown in Figure 4. The center support housing was designed 

with cylindrical end flanges to mate with the forward and aft 

nacelle conical support sections. Rectangular side flanges were 

designed to mate with the rectangular support arms. The center 

support housing could be made from cast steel or fabricated 

from steel as a weldment. The outer contour of the center 

support housing was made from thin rolled steel plate and 

could be attached to the center housing in split sections. Access 

ports were located opposite the rectangular side flanges and at 

other locations to provide access for inspection and 

maintenance. Bulkheads were designed to be located at the 

forward and aft ends to mount the gearbox and to provide 

support for the drive components.

The forward conical support housing was designed to 

support the rotor and drive shaft with the bearing and seal 

housing, shown in Figure 5. It consisted of an aft cylindrical 

bulkhead flange that mated with the forward end of the center 

support housing and a forward bulkhead flange that mated with 

the forward end of the bearing and seal housing near the rotor 

hub. Steel box beams were designed to tie the forward and aft 

steel bulkhead flanges together into an integrated welded 

support structure. The outer skin consisted of rolled steel plates 

with rolled “T-section” rings welded to the inside to provide a 

water-tight pressure hull. This unified structure transferred the 

rotor forces (bending, axial and torsional) from the forward 

conical support housing to the center support housing. The 

design combined strength and rigidity with low weight. The 

design will require further refinement with FEA analysis to 

optimize the structural components.

The aft conical support housing was designed to enclose 

the generator in a water-tight enclosure and to provide access 

on the aft end, shown in Figure 3. It consisted of a forward 

cylindrical flange that mated with the aft end of the center 

support housing and an aft flange that mated with the access 

port on the aft end. The outer skin consisted of rolled steel 

plates with rolled “T-section” rings welded to the inside to 

provide a water-tight pressure hull. The aft access port 

consisted of a hemispherical shell with rolled “T-section” rings 

welded to the inside for support. The design combined strength 

and rigidity with low weight. The design was in the preliminary 

stages and will require further refinement with FEA analysis to 

optimize the structural components. The assembled support 

structures are shown in Figures 7 and 8.

The nacelle was designed for easy system assembly and 

access for maintenance. Access ports are provided in critical 

locations to enable necessary maintenance activities in an 

efficient manner. System build up is performed in center 

housing out fashion starting with the center conical support 

housing and working outward aft and forward. The generator/ 

gearbox assembly is mounted to a sliding rail system permitting 

easy access to these components through removal of the aft 

conical section. Figure 2 illustrates a blow-out of this design.

Maintenance and Life Cycle

The life cycle and preventative maintenance for large PTO 

system components can be comprehensive and can affect the 

overall CoE for a marine hydrokinetic turbine. To estimate an 

accurate CoE for the device, the maintenance and projected 

replacement requirements for the main PTO components were 

estimated based on manufacturer specifications 

/recommendations and design/operation experience with similar 

components used in the marine environment. The following are 

a statement of the data collected for each component. 

Information for the nacelle body was not available though a 

major concern in a marine water environment would be bio 

fouling.

Figure 5. Forward conical support housing. Figure 6. Aft conical support housing.
7

Figure 7. Assembled support structures.

Figure 8. Section through Assembled Support Structures.

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)

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 

Pittsburgh PA. Part # TBD, Price $974.

Component Life: Approximate life 2 years based on user 

experience.

5. Rotor Drive Shaft (1)

Preventative Maintenance: annual inspection, possible 

resurfacing in wear areas.

Component Life: Designed for stresses to be under the 

endurance limit of the material during normal operation. 

Design factor of safety in gust conditions is 2.5. Minimum 

design factor of safety during normal operation is 8. These 

factors of safety are based on overhung load of "DRY" 

rotor weight; pending a more detailed design & system 

analysis including effects of dynamic loading. It should be 

noted the that the drive train has been designed such that 

the shaft and bearings assume all radial loads generated at

the rotor plane in addition to torque and thrust. The 

gearbox is subject to no extraneous rotor loads other than 

torque. The housing containing this shaft, the bearings and 

the associated seals is in one self-contained assembly. This 

assembly can be removed as a completed assembly for 

ease of replacement and efficient maintenance with 

minimal down time. A spare assembly could be maintained 

such that the entire bearing and seal package could be 

removed, inspected and any necessary items replaced or 

repaired with periodic maintenance cycles. With the 

current preliminary bearing selection having an estimated 

life of 16.5 years, this bearing & seal package exchange 

could be done every 2 to 4 years.

6. Generator

Manufacturer: ABB, Model: AMZ500LE10.

Preventive maintenance: Slip ring unit check every year/other 

maintain 5-10 year interval.

Lifecycle: over 10 to 20 years if properly maintained.

7. Transformer

Manufacturer: ABB, Type: 500 kVA Liquid Filled Pad mounted 

Transformer Green-R-Pad FS or EFS.

Preventive maintenance: gauge readings, tank leaks, control 

wiring & circuits, liquid dielectric test, temperature scan, 

insulator cleanliness need to be check once a year.

Lifecycle: with proper maintenance typically greater than 20 

years.

8. Drive Control

Manufacturer: ABB, Model: ACS800-17-0790-7.

Preventive maintenance: Heat sink temperature check & 

cleaning = 6 to 12 months/change cooling fan every 6 

years/capacitor change every 10 years.

Component life: Fan life ~50,000 operating hours/capacitor 

lifespan = 45,000 (~5 years) to 90,000 hours.

Cost & Economic Assessment of the PTO

The lifecycle-cost analysis for the drive train was 

performed by RE Vision Consulting. In order to understand the 

lifecycle cost of the powertrain, the following areas need to be 

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 
8

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 

addressed by use of specialty vessel that is equipped to meet the 

requirements of the task. Total vessel cost was estimated at 

$13million to $16million. A crew of 20 men would be required 

to safely operate the vessel in a tidal race. Other cost elements 

include (1) insurance, (2) replacement part cost, and (3) repair 

personnel cost, (4) facilities and infrastructure.

With the exception of the sea water bearing and seal 

assembly and the supporting structure for the PTO, the balance 

of the drive train uses existing wind turbine technology. To 

estimate the cost of a commercially mature production level, 

results from NREL’s WindPACT study were used and modified 

to take into account the following design differences between 

wind and water turbine designs:

1. Lower rotor shaft rpm resulting in increased gearbox cost

2. Difference in rotor designs

3. Design redundancies in the drive train to minimize the 

number of annual repairs for the water turbine

4. Sealing requirements

The following figure shows the relative cost comparison 

between a wind-power drive-train, an equivalent tidal turbine 

drive-train and the redundant tidal-turbine drive-train.

Figure 9: Capital cost of different drive-train configurations

It is important to understand that the design of a powertrain 

cannot be carried out in isolation from the system design.

Lifecycle cost data from all systems were used to establish a 

comprehensive lifecycle cost model. The resulting data was 

subsequently used to optimize the intervention intervals for the 

powertrain. While actual cost of electricity data for this project 

has not been released yet, the following illustration shows the 

contribution of individual cost centers to the total system cost 

(O&M costs are not shown in this figure). It shows that the 

powertrain’s contribution to total cost will likely be significant 

at commercial scale.

Acknowledgments

The authors would like to acknowledge the support of the 

US Dept. of Energy and Michael Reed, Program Manager, for 

support of this research design effort. DOE Research contracts 

#775396.

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