SERITTP-253-3616 

UC Category: 64 

DE90000304 

Conceptual Design of an 

Open-Cycle Ocean Thermal 

Energy Conversion Net 

Power-Producing Experiment 

(OC-OTEC NPPE) 

A cooperative effott of 

Solar Energy Research Institute 

Pacific International Center for 

High Technology Research 

Argonne National Laboratory 

D. Bharathan 

H. J. Green 

H. F. Link 

8. K. Parsons 

J. M. Parsons 

F. Zangrando 

Prepared under Field Work Proposal OE91 

Solar Energy Research lnstitute 

A Division of Midwest Research lnstitute 

161 7 Cole Boulevard 

Golden. Colorado 80401 -3393 

Prepared for the 

U.S. Department of Energy 

Contract No. DE-AC02-83CH10093 
NOTICE 

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PREFACE 

The mission of the Ocean Energy Technology (OET) program of the U.S. 

Department of Energy (DOE) is to 

develop ocean energy technology to a point where the 

commercial sector can make competent assessments of whether 

applications of the technology are viable energy conversion 

alternatives, or supplements, to systems presently in use (U.S. 

DOE 1985). 

Toward this mission, the program is aimed to advance the technology over 

the next several years. To address research and developmental issues, focus 

is placed on resolving the technical uncertainties related to near-shore 

ocean thermal energy conversion (OTEC) systems with capacities ranging 

from 2 to 5 MWe. Efforts will include 

Emphasizing power conversion research by improving heat- and masstransfer process, reducing component size and water flow requirements, 

and using less costly materials and components 

Increasing experimental verification and testing activities needed to 

confirm the feasibility and performance boundaries of critical 

components. 

In 1987, a cooperative effort by DOE, the State of Hawaii, and the Pacific 

International Center for High Technology Research (PICHTR) upgraded the 

seawater supply system at the Seacoast Test Facility (STF) at the Natural 

Energy Laboratory of Hawaii (NELH). This upgrade expands the OET 

program's capability to experimentally investigate critical components and 

precommercial OTEC systems. 

The recent tests of heat exchangers at the Heat- and Mass-Transfer 

Scoping Test Apparatus (HMTSTA) have confirmed their performance 

projections using seawater. Research has led to the conceptual design of 

the Heat- and Mass-Transfer Experimental Apparatus and the Net PowerProducing Experiment (HMTEA/NPPE), which extend the testing capability 

sixfold from HMTSTA and provide for full system tests to be carried out. 

The design was pursued cooperatively by the Argonne National Laboratory 

(ANL), PICHTR, and the Solar Energy Research Institute (SERI). 

This design, based on an extensive array of analytical tools and 

experimental data, permits the OET program to proceed with confidence in 

development and operation of the HMTEAINPPE. 
The HMTEAJNPPE system is built around a single, vertical-axis, mixed-flow turbine 

rotor 2.9 m (9.5 ft) in diameter, as is illustrated on the opposite page. The rotor is 

supported by a vacuum vessel 7.3 m (24 ft) in diameter and 7.6 m (25 ft) high which is the 

structural backbone of the system. The electrical generator is an air-cooled, salientpole, synchronous generator located above the turbine assembly and outside the vacuum 

enclosure. 

Steam is produced in an annular flash evaporator at the periphery of the vacuum vessel, 

which consists of 12 vertical spouts, each 0.24 m (10 in.) in diameter. The steam flows 

up from the evaporator, through a mist eliminator located at the turbine inlet flange, and 

enters the turbine radially inward. The steam exits the turbine axially in the center of 

the vessel. A conical exhaust diffuser with center body is used to recover most of the 

kinetic energy of the steam discharged from the turbine. The diffuser is 3.65 m (12 ft) 

long and provides an annular entry to the first condenser stage. 

The condenser is a direct-contact, structured-packing condenser composed of two coaxial 

stages. The first stage fills the outer annular space of the condenser assembly, between 

the diameters of 4.11 m (13.5 ft) and 2.43 m (8 ft), and it receives steam from the 

turbine diffuser. It operates in cocurrent flow, with seawater and steam flowing 

vertically downward through 0.61-m- (2-ft)-high packing material. The second stage is in 

the center of the condenser assembly and operates in countercurrent flow to condense 

the steam left over from the first stage, so that less pumping is required of the vacuum 

vent system. The countercurrent stage contains the same packing material as the first 

stage. It is smaller and longer, with an overall packing height of 0.91 m (3 ft). Cold 

seawater is supplied from below the condenser through a pipe that rises above the 

condensers at the center of the structure and that delivers all the seawater for both 

stages. The distribution manifold provides cold seawater to both stages through a drip 

tray for the countercurrent stage and a series of radial pipes for the cocurrent stage. 

The noncondensable gases liberated from the seawater streams and a small amount of 

uncondensed steam are compressed and exhausted using a vacuum vent system. 

A surface condenser outside the vacuum enclosure condenses 10% of the steam flow in 

order to produce desalinated water as a by-product of power generation. 

At the end of the conceptual design, the components have been specified in sufficient 

detail that overall system performance may be predicted. Analysis projects a gross 

power of 213 kWe and a net power of 84 kWe for the system at the baseline design 

condition. The sensitivity of system performance to uncertainties in key variables 

results in a standard deviation of 5 kWe for the net power. Considering these 

uncertainties, a minimum net power from the NPPE is projected to be no less than 

67 kWe at design. 

The small fraction of gross power available as net for the NPPE makes design and 

implementation of the system challenging. Replacement of existing onshore seawater 

supply lines is required. Less-than-projected performance by the offshore pipes, vacuum 

system, and seawater disposal trench under development can result in a net power of 

around 50 kWe. However, at this stage of the design, projections indicate that the goal 

of the project will be met. 

Project cost is estimated at $15 million over the next three fiscal years for the design 

and development. The HMTEA will be operational in mid-1991 and the NPPE in mid1992. 
HMTEAJNPPE 

Open-Cycle 

OTEC System 

Spout d* 

Photograph of vacuum structure showing internal 

arrangement of HMTEAINPPE system components 

vii 
TABLE OF CONTENTS 

Page 

1.0 Introduction ...................................................... 1 

1.1 Background on Open-Cycle Development ........................... 2 

1.1 . 1 Background on the Seawater Systems ...................... 3 

1.1.2 Summary ............................................... 3 

1.2 HMTEAJNPPE Overview ......................................... 3 

1.2.1 Summary ............................................... 3 

1.2.2 Task Areas ............................................. 4 

1.2.3 Technical Approach ..................................... 5 

1.3 Management Approach ........................................... 5 

1.3.1 Organizational Roles and Responsibilities ................... 5 

1.3.2 Planning .............................................. 6 

1.3.3 Project Review and Control .............................. 6 

1.4 Experiment Constraints ........................................... 10 

2.0 System Integration ................................................... 11 

2.1 Goals and Objectives ............................................ 11 

2.2 Design Requirements ............................................ I1 

2.3 System Analysis ................................................. 13 

2.3.1 Analysis Method ........................................ 13 

2.3.2 Analysis Results ........................................ 14 

2.3.3 Sensitivity Analysis ...................................... 16 

2.3.4 Code Benchmark Comparisons ............................ 20 

2.3.5 System Control ......................................... 21 

2.4 Physical Integration ............................................. 26 

2.4.1 Integration Options ...................................... 26 

2.4.2 Selection Criteria and Rationale .......................... 26 

2.4.3 Configuration Options ................................... 27 

2.4.4 The Integrated NPPE System Design ....................... 30 

2.4.5 Site Plan ............................................... 38 

2.4.6 System Interfaces ....................................... 38 

2.5 Cost Projections ................................................ 38 

2.5.1 Design and Construction ................................. 42 

2.5.2 Technical Management ................................... 42 

2.6 Remaining Issues and Future Work ................................. 43 

2.6.1 Technical Issues ......................................... 43 

2.6.2 Preliminary Design ...................................... 46 

2.6.3 Final Design ............................................ 47 

3.0 Turbine and Structures ................................................ 49 

3.1 Development of the Turbine ...................................... 49 

3.1.1 Goal and Objective ...................................... 49 

3.1.2 Approach ............................................ 49 

3.1.3 Options Considered and Eliminated ........................ 49 

3.1.3.1 Existing Hardware vs . Existing Technology ......... 49 

3.1.3.2 Horizontal vs . Vertical Axis ..................... 50 

3.1.4 Other Major Considerations ............................... 51 

3.1.4.1 Radial-Inf low vs . Axial Configuration ............. 51 
TABLE OF CONTENTS (Continued) 

Page 

3.1.4.2 Modified Steam Conditions ...................... 52 

3.1.4.3 Speed Control ................................. 52 

3.1.4.4 Variable-Position Stator Vanes ................... 53 

3.1.5 Design Performance and Uncertainties ..................... 53 

3.1.6 Design Specifications .................................... 55 

3.1.7 Layouts ................................................ 55 

3.1.8 Interfaces .............................................. 55 

3.1.9 Schedule ............................................... 58 

3.1.10 Cost and Uncertainties ................................... 58 

3.2 Structures ...................................................... 58 

3.2.1 Goal and Objective ...................................... 58 

3.2.2 Approach .............................................. 60 

3.2.3 Options Considered and Eliminated ........................ 61 

3.2.3.1 Dome vs . Telephone Configuration ................ 61 

3.2.3.2 Materials for Vacuum Vessel Walls ................ 62 

3.2.4 Other Major Considerations ............................... 65 

3.2.4.1 Position of the Condenser ....................... 65 

3.2.4.2 Diameter of the Vacuum Vessel .................. 65 

3.2.4.3 Height of the Vacuum Vessel ..................... 65 

3.2.5 Design Performance and Uncertainties ..................... 66 

3.2.6 Design Specifications .................................... 67 

3.2.6.1 Vacuum Vessel ................................. 67 

3.2.6.2 Control and Auxiliary Buildings .................. 69 

3.2.6.3 Crane ........................................ 69 

3.2.7 Layouts ................................................ 69 

3.2.8 Interfaces .............................................. 70 

3.2.9 Schedule ............................................... 71 

3.2.10 Cost and Uncertainties ................................... 71 

4.0 Seawater Systems .................................................... 73 

Objective ...................................................... 73 

Goal .................................................... 73 

Approach ...................................................... 73 

4.3.1 Design Conditions ....................................... 73 

4.3.2 Pump Configuration ..................................... 74 

4.3.3 Pump Bids .............................................. 76 

4.3.4 Electric Motors ......................................... 78 

4.3.5 Control of Flow Rate .................................... 81 

4.3.6 Supply and Discharge Pipes ............................... 81 

4.3.7 Concrete Sump ......................................... 82 

4.3.8 Disposal Trench ......................................... 83 

4.3.9 Seawater System Tests ................................... 84 

4.3.10 Environmental Issues .................................... 85 

4.3.11 Plant Safety ............................................ 85 

Layout ....................................................... 85 

Performance ................................................... 87 

Specifications .................................................. 88 

Interfaces ...................................................... 88 

Schedule ....................................................... 88 
TABLE OF CONTENTS (Continued) 

Page 

4.9 Cost Estimate ................................................ 90 

5.0 Heat Exchangers ..................................................... 95 

5.1 Objective and Goals ............................................. 95 

5.1.1 Composition of the Heat Exchanger Task Team 

and Documentation ...................................... 95 

5.2 Approach to Design .............................................. 97 

5.2.1 Baseline and Nominal Operating Conditions ................. 97 

5.2.2 Major Geometric and Operational Considerations ............ 98 

5.2.3 Secondary Heat Exchanger for Desalinated Water 

Production ............................................ 99 

5.2.4 Predeaerators .......................................... 101 

5.3 Conceptual Design of Heat Exchanger Components .................. 102 

5.3.1 Evaporator ............................................ 103 

5.3.1.1 Evaporator Planform Area and Loading ............ 103 

5.3.1.2 Evaporator Spouts .............................. 106 

5.3.1.3 Evaporator Water Supply and Distribution ......... 106 

5.3.1.4 Flow Variation in the Evaporator 

Distributor .................................... 107 . 5.3.1.5 Evaporator Discharge Plenum .................... 107 

5.3.2 Mist Eliminator ......................................... 108 

5.3.3 Direct-Contact Condenser (DCC) .......................... 108 

5.3.3.1 DCC Planf orm Area and Required Loadings ........ 110 

5.3.3.2 Structured Packing and Condenser Heights ......... 111 

5.3.3.3 Water Distributors for the DCC .................. 112 

5.3.3.4 Control of Water Flow .......................... 113 

5.3.3.5 Discharge Plenum of the Direct-Contact 

Condenser ................................... 114 

5.3.3.6 Gas Vent Pipe ................................. 115 

5.3.4 Hydraulic Losses of Main Heat Exchangers .................. 115 

5.3.5 Surface Condenser for Desalinated Water Production ......... 116 

5.4 Design Performance and Uncertainties ............................. 116 

5.5 Design Specifications ............................................ 117 

5.6 Interfaces ...................................................... 117 

5.7 Schedule .................................................. 119 

5.8 Cost Estimates ............................................... 119 

6.0 Ancillary Systems .................................................... 125 

6.1 Objective ................................................... 125 

6.2 Goal ........................................................... 125 

6.3 Vacuum Compression System ..................................... 125 

6.3.1 Design Approach ........................................ 126 

6.3.2 Inlet Conditions ......................................... 126 

6.3.3 Compressors ............................................ 128 

6.3.4 Gas Condensers and Coolers .............................. 130 

6.3.5 Compressor Drivers ..................................... 132 

6.3.6 Other Vacuum Compressor Support Equipment .............. 133 

6.3.7 Conceptual Design Performance Prediction ................. 133 
TABLE OF CONTENTS (Concluded) 

Page 

6.3.8 Preliminary Design Contract .............................. 135 

6.3.9 Remaining Issues and Uncertainties ........................ 135 

6.4 Electrical System ............................................... 135 

6.4.1 Design Approach ........................................ 135 

6.4.2 System Requirements and Options ......................... 135 

6.4.3 Cost Estimates ........................................ 137 

6.4.4 Remaining Issues and Uncertainties ........................ 137 

6.5 Instrumentation and Control Systems .............................. 137 

6.5.1 Design Approach ........................................ 137 

6.5.2 Configuration ........................................... 138 

6.5.3 Cost Estimate .................... .. ................... 139 

6.5.4 Remaining Issues ........................................ 139 

6.6 Miscellaneous Ancillary Systems .................................. 140 

6.6.1 Design Approach ........................................ 140 

6.6.2 Cooling-Water Circulation ................................ 140 

6.6.3 Desalinated-Water System ................................ 141 

6.6.4 Air-conditioning ........................................ 141 

6.6.5 Other Miscellaneous Items ................................ 141 

6.7 Summary ....................................................... 141 

7.0 Concluding Remarks .................................................. 145 

8.0 References .......................................................... 147 

8.1 OpenLiterature ................................................. 147 

8.2 Internal Documents .............................................. 149 

Selected Distribution List .................................................. 153 

xii 
LIST OF FIGURES 

Page 

Work breakdown structure for the development of HMTEAJNPPE .......... 7 

Research activities and plan for the development of HMTEAJNPPE ........ 8 

Schedule for design and developmental activities for HMTEAJNPPE ........ 9 

Variation of gross power vs . steam flow for the net power-producing 

experiment (NPPE) .................................................. 15 

Variations of HMTEAJNPPE system net power with resource-seawater 

flow rates .......................................................... 16 

Sketches of lltelephonelv and "dome" configurations for HMTEAJNPPE ...... 27 

General arrangement of components in vvtelephonerw configuration 

for HMTEAINPPE ................................................... 28 

General arrangement of components in "domevv configuration for 

HMTEAINPPE ................................................... 30 

Engineering sketch of HMTEAJNPPE ................................... 31 

Process flow diagram for HMTEAINPPE. with critical interfaces 

identified .......................................................... 32 

Photograph of vacuum structure showing internal arrangement of 

HMTEAINPPE system components ..................................... 35 

Site plan for HMTEAINPPE at the DOE Seacoast Test Facility. 

Natural Energy Laboratory of Hawaii. Kailua-Kona ...................... 39 

Approach to turbine development ...................................... 50 

Operating conditions for system and turbine ............................ 53 

Turbine off -design performance characteristics ......................... 54 

Design concept for the turbinelgenerator ............................... 56 

Physical interface between turbine and structure ........................ 57 

Diagram of principal interfaces. NPPE turbine .......................... 59 

A partial representation of Dillingham design of the HMTEA .............. 61 

Steel wall option for vacuum vessel .................................... 63 

Platform layout (a) and penetration layout (b) of vacuum vessel ............ 66 

... 

Xlll 
LIST OF FIGURES (Concluded) 

Page 

Layout of control building and associated structures ..................... 68 

Vertical barrel or can pump ........................................... 76 

Vertical wet pit pump installation in the pump station .................... 77 

Sample warm-seawater pump curves ................................... 79 

Sample cold-seawater pump curves .................................... 80 

Sample cold-seawater pump curves .................................... 80 

Concrete sump and 1.22-m (48-in) discharge pipe with spool piece 

(R.M.Towill1989) ................................................... 84 

Warm-seawater system heads and elevations. draft. 9/22/89 ............... 86 

Cold-seawater system heads and elevations. draft. 1 1/08/89 ............... 87 

Seawater systems interface diagram. 12/6/89 ........................... 91 

Heat exchanger interface diagram ..................................... 96 

Plan view of heat exchangers and evaporator distributor .................. 104 

Elevation view of evaporator ......................................... 105 

Influence of liquid loading on seawater spout evaporator .................. 106 

Elevation view of direct-contact condenser ............................. 109 

Model parametric runs for NPPE. packing and freefall heights ............. 112 

Elemental section of test unit ......................................... 117 

Overall view of condenser test unit .................................... 118 

Conceptual design of vacuum compressor system ........................ 126 

Pressure-volume diagram for centrifugal and positive-displacement 

types of machines ................................................... 134 

xiv 
LIST OF TABLES 

Page 

2-1 HMTEAINPPE Seawater System Design Conditions as of 

September 14. 1989 .................................................. 12 

2-2 Power Budget Distribution and Design Projections ....................... 14 

2-3 Nominal Design Conditions for the NPPE System 

(Projections at Nov . 1989) ............................................ 17 

2-4 NPPE System Net Power Sensitivities .................................. 19 

2-5 Systems Code Benchmark Comparison of NPPE Design Conditions ......... 22 

2-6 HMTEAJNPPE System Specifications .................................. 33 

2-7 List of Critical System Interfaces ..................................... 40 

2-8 HMTEAJNPPE Project Costs at Nov . 1989 .............................. 44 

3-1 Preliminary Design Evaluation for Turbine .............................. 51 

3-2 Steam Inlet and Outlet Conditions for the NPPE Turbine .................. 52 

3-3 Schedule for NPPE Turbine ........................................... 60 

3-4 NPPE Turbine Costs ................................................. 60 

3-5 Schedule for NPPE Structures Task .................................... 71 

3-6 NPPE Structure Costs ............................................... 71 

4-1 NPPE Seawater Systems Performance Summary (as of 11-15-89) ........... 89 

4-2 Seawater Systems Specifications as of 1 1- 15-89 ......................... 90 

4-3 Seawater Systems Schedule (as of 1 1.15.89) ............................. 91 

4-4 Cost Estimate for the HMTEAINPPE Seawater Systems .................. 92 

5-1 Options for Secondary Condensers.. Desalinated Water Production ......... 100 

5-2 Performance Predictions at Nominal Conditions (Summary) ............... 119 

5-3 Heat Exchangers: Specifications and Related Parameters ................ 120 

5-4 Projected Schedule for Heat Exchangers ................................ 122 

5-5 Breakdown of Heat Exchanger Costs ................................... 123 

6-1 Key Performance Parameters for a VCS ................................ 127 
LIST OF TABLES (Concluded) 

Page 

Basis for Defining the Requirements for the VCS Design .................. 127 

Inlet NPPE Operating Requirements for the Preliminary VCS .............. 128 

Design Contract 

Compression System Power Projection Using Existing PositiveDisplacement Compressor Hardware at Standard Pressure Ratios .......... 129 

Compression System Power Projection Using Existing PositiveDisplacement Compressor Hardware at Lower-Than-Standard 

Pressure Ratios .................................................... 130 

Summary of Gas Condenser/Cooler Types Considered for the 

Precooler and Intercooler Designs of the HMTEA/NPPE VCS .............. 131 

Cost Estimate for Ancillary Electric System ............................ 137 

Instrumentation and Controls Needed for the HMTEAINPPE .............. 138 

Power Consumption Estimate for Ancillary Systems ...................... 142 

Cost Estimate for Ancillary Systems ................................... 142 

Summary Schedule for Ancillary Systems ............................... 142 
CHAPTER 1.0 

INTRODUCTION 

Ocean thermal energy conversion (OTEC) technology is based on the principle that 

energy can be extracted from two reservoirs if they are different temperatures. A 

temperature difference of about 20°C can be exploited effectively to produce usable 

energy. Temperature differences of this magnitude exist between ocean waters at the 

surface and at depths up to 1000 m in many areas of the world, in tropical latitudes 

between 24 deg north and south of the equator. Here, surface water temperatures typically range from 24°C to 29OC, and temperatures at a depth of 1000 m typically range 

from 4°C to 6OC. This constitutes a vast, renewable resource for baseload power generation. 

The U.S. Department of Energy (DOE) has concentrated on two OTEC power cycles for 

converting this thermal energy to electrical energy, namely, closed cycle and open 

cycle. In a closed-cycle system, warm seawater is used to vaporize a working fluid, such 

as ammonia, flowing through an evaporator. The vapor expands at moderate pressures 

and turns a turbine. The vapor is condensed in a condenser using cold seawater pumped 

from the ocean's depth through a cold-water pipe. The condensed working fluid is recirculated to the evaporator to repeat the cycle. 

In an open-cycle system, steam generated from the warm seawater is the working fluid. 

A small portion of the warm seawater is flash-evaporated in a vacuum chamber at an 

absolute pressure of about 2.4 kPa (0.34 psia) to produce steam. The steam expands 

through a low-pressure turbine coupled to a generator to produce electricity. The steam 

exiting the turbine is condensed using the cold seawater in a direct-contact condenser. 

If a surface condenser is used, the condensed steam provides desalinated water as well. 

Open-cycle research is based on the projected advantages of water production and of 

completely or partially eliminating conventional heat exchangers in the cycle and the 

penalties of performance and cost associated with them. Whether the full potential for 

this approach can be realized depends on achieving acceptable levels of power-conversion 1 efficiency and on reducing parasitic losses in the highly coupled process in systems of 

sufficiently large scale. 

Analytical efforts coupled with experimental verification on components have indicated 

that open-cycle systems appear promising. These results lead to experimentally investigating the feasibility of this to validate the analytical projections on a system level. 

Combining experimental efforts for the open-cycle heat exchangers with those of the 

turbine subsystem will result in a validated technical data base for the open-cycle power 

system. System-level experiments will be carried out through design, construction, and 

operation of the Heat- and Mass-Transfer Experimental Apparatus (HMTEA) and the net 

power-producing experiment (NPPE). 

The purpose of the present investigation is to determine the interaction among the 

various components, to determine the net-to-gross power output, and to validate and 

refine analytical models. Practical experience gained in these experiments will provide a 

basis for substantiating and modifying the assumptions used in the analyses. A second 

outcome of the research will be the verification of the technical feasibility of the opencycle (OC) OTEC process, over an extended period of time, at a precommercial scale. 
I. 1 Background on Open-Cycle Development 

In 1979, Westinghouse Corporation, with DOE funding, completed design studies for a 

100-MW OC-OTEC plant (Westinghouse 1979). A design was developed for a floating 

system &at used a large vertical-axis turbine, a channel flash evaporator, and a surface 

condenser enclosed in a reinforced concrete vacuum vessel. This design described a castconcrete containment structure that combined multiple structural functions with an 

axisymmetric layout for the major system components. The study concluded that an 

open-cycle plant could be a cost-effective alternative to the closed-cycle designs. It 

identified flash evaporation, direct-contact condensation, and use of innovative materials 

for the turbine and structural elements as ways to further reduce costs significantly. 

A study in 1985 completed by a team from the Florida Solar Energy Center (FSEC) and 

Creare R&D, Inc., explored the economic and technical viability of OC-OTEC for 

producing electricity and desalinated water based on hardware reflecting the current 

state of the art (Block and Valenzuela 1985). FSEC and Creare concluded that such 

plants could deliver desalinated water and electricity at a competitive cost to many 

island markets, with plant sizes of less than 20 MWe and, in many cases, with plants in 

the 2-10 MWe size range. 

The FY 1985-1989 federal Ocean Energy Technology (OET) Multiyear Program Plan 

(issued in December 1985) defined the long-range mission for the program as follows: 

to develop ocean energy technology to a point where the commercial sector 

can make competent assessments of whether applications of the technology 

are viable energy conversion alternatives, or supplements, to systems 

presently in use (US. DOE 1985). 

This mission would be accomplished by researching components and modeling systems. 

A fresh-water direct-contact heat- and mass-transfer laboratory was designed and built 

at the Solar Energy Research Institute (SERI) to study and improve methods for transferring heat and mass under small driving forces at low operating pressures. A number of 

evaporator geometries were tested in this facility. These tests led to the definition and 

experimental evaluation of a vertical-spout evaporator, which has a thermal effectiveness ranging from 90% to 97% with a spout head loss of approximately 0.7 m (Bharathan 

and Penney 1983). 

Extensive experiments were carried out on various configurations of contacting media 

suitable for condensing steam at the low pressures applicable for OC-OTEC. These 

experiments identified condensers using structured packing to yield an effectiveness in 

the 86% to 90% range with minimal vapor pressure loss and a water head loss of less than 

2.0 m (Bharathan, Parsons, and Althof 1988). 

The data from these experiments were used to validate detailed analytical models describing the performance of the direct-contact condensers (Bharathan, Parsons, and 

Althof 1988). The development of this analytical capability and the definition of hardware concepts that achieve the performance potential inherent in the direct-contact 

processes were significant steps in the OET program. 

Most recent research efforts in the OET program were to scale up the experimental 

apparatus to a thermal capacity of over 1 MWt using seawater as the process fluid. This 

apparatus, called the Heat- and Mass-Transfer Scoping Test Apparatus (HMTSTA), was 

designed and built cooperatively by the Argonne National Laboratory (ANL) and SERI and 
is installed at the Seacoast Test Facility (STF) at the Natural Energy Laboratory of 

Hawaii (NELH) (Parsons et al. 1989). 

Tests in the HMTSTA confirmed the heat- and mass-transfer performance of spout 

evaporators and structured packing condensers observed in fresh-water experiments and 

predicted by analytical models (Zangrando et al. 1989). Data on the performance of 

surface condensers and on seawater deaeration also were obtained in HMTSTA experiments. 

Over the past decade, sufficient data and operational experience have been gathered to 

permit the OET program to undertake the next step with confidence, namely, designing a 

larger system, the HMTEAINPPE. 

A parallel set of activities toward developing a suitable turbine for the NPPE began in 

October 1987 with a workshop held at the Pacific International Center for High Technology Research (PICHTR). The outcome of the workshop led to two approaches, namely, 

using existing hardware and pursuing existing technology to develop the turbine. PICHTR 

investigated a variety of existing rotor stages for their potential use. In March 1988 

SERI released a Request for Proposal to turbine manufacturers for developing a turbine. 

Preliminary designs were pursued using both approaches; they were evaluated in July 

1989. The evaluations led to the choice of a newly designed turbine, primarily to minimize the technical risk. The early start of these activities in turbine development resulted in the most complete design to date of any of the major components for the NPPE. 

1.1.1 Background on the Seawater Systems 

In 1987, the warm and cold resource seawater at the STF at the NELH became more \ available. In a joint effort by DOE, the State of Hawaii, and PICHTR, a state-of-the-art 

cold-water pipe was installed. This polyethylene pipe is 1 m in diameter and 2060 m long 

and is deployed to a water depth of 675 m, with the deeper section in an inverted buoyant 

catenary configuration (Lewis, Van Ryzin, and Vega 1988). Along with an upgraded 

warm-seawater intake systefn, the facility an now deliver volume flows of cold and 

3 

5 warm seawater of 0.41 m /s and 0.61 m /s, respectively, to the OTEC facilities. 

Another 0.43 m /s of cold water is delivered to other NELH experiments, primarily in 

mariculture A cooperative agreement between DOE and the State of Hawaii stipulates 3 that 0.41 m /s of the total cold water will be available to the continuing research activities in the OET program. 

1.1.2 Summary 

The sequence of steps followed over the past decade in developing the OC-OTEC technology and installing a seawater system at the NELH has led to the conceptual design of 

the HMTEAINPPE described in the remainder of this report. This design, based on an 

extensive array of analytical tools and experimental data, permits the OET program to 

proceed with confidence to a net-power-generating system for performance studies and 

verification. 

1.2 HMTEAINPPE Overview 

1.2.1 Summary 

The design of the HMTEAINPPE builds on performance data obtained on heat exchangers 

using seawater at the HMTSTA (Zangrando et al. 1989), turbine design studies (Mechanical Technology Inc. 1989; Fong, Masutani, and Neil1 19891, and HMTEA design studies by 
industry teams (Dillingham 1987a; PICHTR 1988). The apparatus will be designed and 

built in phases. The HMTEA will permit a sixfold scale-up from the HMTSTA of heatand mass-transfer hardware to validate predictive codes at a scale consistent with 

resource water availability. The HMTEA will also provide data that will facilitate effective operation of the NPPE with minimum technical risk for producing net power. The 

NPPE will be complete when a turbinelgenerator unit is integrated into the HMTEA. 

The experimental testing of this system will meet the OET program goals when net 

power is produced at a design goal of 40 kW The data gathered with the HMTEAJNPPE 

will augment the technical data base to &;date system performance predictive codes 

and to understand the system component interactions and stability and control characteristics of a power-producing system. The operational experience will generate insights 

into the dynamic stability; transient response; and the long-term reliability, availability, 

performance, and maintainability of the system. 

1.2.2 Task Areas 

The design of the HMTEAJNPPE is organized into five major technical areas for 

managing the design effort: 

Systems integration 

TurbineJstructures 

Seawater systems 

Heat exchangers 

Ancillary systems. 

The systems integration task provides overall guidance to the HMTEAJNPPE development. Setting directions and goals, planning, and managing all aspects of the development fall under this task. 

The turbine/structures task carries forward the designs for the turbine and structures 

within the general framework established by the design team. Development of all 

support facilities is included in this task. 

The seawater systems task addresses issues related to the supply and discharge of the 

seawater to and from the facility. Other areas under this task include safety, permits, 

and environmental issues. 

' The heat exchanger task addresses the design and development of the direct-contact and 

surface heat exchangers. Other areas included under this task are predeaeration and 

activities related to heat- and mass-transfer tests. 

The ancillary systems task addresses issues related to the development of all remaining 

components and systems. Specifically these areas include the development of the 

vacuum system, electrical systems, miscellaneous cooling systems, instrumentation and 

control, and data-acquisition systems. 

For the conceptual design, the systems integration analysis treats the other four tasks 

representing the major subsystems as "black boxes" with a projected practical performance. This analysis defines the design and operating parameters to achieve the 

required system performance and establishes the design interfaces among the components. The details of the analyses, designs, and specifications that describe the systems 

integration effort and each of the major subsystems are presented in subsequent chapters 

of this report. The system analysis and physical integration design presented in 
Chapter 2.0 establishes the design parameters for each of the components and is a 

necessary precursor to the detailed design work. Descriptions of the design effort 

carried out in the other four task areas follow in the subsequent chapters. 

1.2.3 Technical Approach 

The technical approach to developing the HMTEAJNPPE is to complete a three-phase 

design--conceptual, preliminary, and engineering--before constructing the experiment. 

In each phase of the design, all aspects of the system should be addressed, with increasing detail as the design progresses. The products of all phases of the design will be in the 

form of drawings and specifications, performance analyses and projections, cost estimates, schedules, and status reports of progress. 

The following defines the differences in the products for each of these cycles. 

Conceptual Design. The purpose of the conceptual design is to identify and evaluate an 

experimental concept that can most effectively meet the research goal for the project. 

In the conceptual design, investigators must emphasize the performance analysis using 

the best tools available to identify the mix and arrangement of subsystems and components that can best achieve the stated goal. The drawings developed are engineering 

\ sketches to define the overall configuration and arrangement of key components and 

subassemblies. Cost estimates will be provided to assess that the experiment can be 

completed within the overall program budget. At this time these costs must be considered rough-order-of-magnitude estimates. 

At the end of the conceptual design, the type, specifications, and arrangement of the 

major subassemblies to achieve the experimental objectives will be established. 

Preliminary Design. The purpose of the preliminary design is to establish the engineering 

and economic feasibility of the experimental concept (established earlier) by identifying 

the engineering design and reconfirming the performance projections of the concept. 

The mix between the drawings and analysis shifts toward design and cost activities; the 

level of analytical efforts increases in an attempt to reduce the technical and cost 

t uncertainties to less than 25%. This implies that the analytical models and assumptions 

need to be rigorously justified, and the drawings need to provide sufficient details that 

engineering realities are reflected in the cost and performance projections. 

At the end of the preliminary design, the system design and detailed subassembly specifications will be established. 

Engineering Design. The purpose of the final engineering design is to produce the design 

and documentation necessary to build and install the experiment. Thus, the emphasis 

shifts almost entirely to the design of subassemblies and components. The only analysis 

required is to evaluate such changes in subsystem performance arising out of detailed 

engineering design and to develop the detailed test program. 

1.3 Management Approach 

1.3.1 Organizational Roles and Responsibilities 

The management approach for the HMTEAINPPE project is built around implementing a 

cooperative team effort involving SERI, ANL, and PICHTR working under DOE 

direction. The management process was initiated in a planning session involving the first 
working group of the three participating organizations on February 22-23, 1989 (Presentation Packages 1989). In this session and the plan that resulted from it, five task teams 

were formed that included members from each laboratory and a representative from 

DOE. A task leader was identified for each area. SERI was given the lead role for 

HMTEAJNPPE system development in the conceptual design; ANL and PICHTR were 

given supporting roles. All organizations will participate in the design and implementation. PICHTR will assume lead responsibility in the later design stages and implementation. 

1.3.2 Planning 

The first working group reviewed and clarified the HMTEAINPPE plans that were developed, established teams for completing the project, defined operational and review 

procedures, and established a baseline configuration for the HMTEAINPPE that would 

form the point of departure for all future design work. Out of this planning activity, a 

system development management plan (First Working Group 1989) was developed on the 

basis that 

a The NPPE would be ready for testing in the first quarter of 1993. 

The needed resources in terms of labor and dollars were available. 

a The defined 'baseline" configuration would remain the concept from which further 

variations would occur. 

The management plan organizes the project according to the system skeleton work 

breakdown structure shown in Figure 1-1. This structure follows a detailed structure 

developed specifically for OTEC systems (Value Engineering 1979). 

The HMTEAJNPPE project is scheduled in the time frame indicated on the OET program 

activities chart shown in Figure 1-2. 

A detailed Gantt chart for the project is shown in Figure 1-3. Major milestones in the 

project are completing the three stages of design, the construction and installation of the 

system, and the performance tests in the HMTEA and the NPPE. The items on the 

critical path toward the project completion are the development of the turbine for the 

NPPE and the structure and vacuum exhaust systems for the HMTEA. Further assessments of the schedule during the conceptual design stages indicate that with an accelerated effort, the HMTEA will be operational in mid-1991 and the NPPE in mid-1992, 

provided the current levels of effort can be maintained. 

The NPPE systems team is responsible for reviewing and updating all elements of the 

management plan. 

1.3.3 Project Review and Control 

In addition to the normal program control exercised by DOE through the annual operating 

plan and field work proposals, the HMTEAINPPE project will incorporate formal review 

procedures that will provide for consulting, evaluation, and recommendation services to 

each of the task areas. 

The individual task teams are responsible for resolving technical issues. If an issue 

cannot be resolved at the task level, it will be referred to the systems team for resolution. If resolution does not occur at that level, it will be carried forward to DOE for 

final resolution. 
At the end of each phase, DOE, SERI, ANL, and PICHTR, with provision made for independent evaluations by selected expert(s) from outside the program, will review the 

project before it moves to the next development phase. The form for this review will be 

a formal presentation of project results, status, and unresolved issues documented in a 

technical report similar to this conceptual design report. 

The individual task leader is the communication facilitator, keeping the systems integrations task leader informed through weekly highlights and direct and special communication. In writing to the systems integration task leader, the individual task leader documents specific work assignments, specific conclusions or recommendations made by the 

team, unresolvable issues, or factors affecting other tasks. The individual task leader is 

reponsible for maintaining communication and information flow among his or her task 

team members. 

The primary task control vehicle is a task plan produced by the task leader and approved 

by the systems integration team. The plan will communicate task objectives, strategy, 

and tactics. The technical milestones for the organization's annual plan will be incorporated into the project plan and will be meaningful measurements of the percentage of 

completion of a given task. The milestones will be used by the systems integration task 

leader to evaluate progress against the goal and to recommend corrective or preventive 

action as required. 

NPPE/OC-OTEC 

System Development 

k3.5 Ancillary Subsystems 

, 

Figure 1-1. Work breakdown structure 

for the development of MTEA/NPPE 

t 3.6 Seawater Supply 

Subsystem 

General 

Management 

1 .O 

L3.7 Total System 

Acceptance 

- 1 .I Design Phase 

Management 

- 1 2 ACQIConstruction 

and Deployment 

Phase Management 

Design 

2.0 

- 2.1 Conceptual 

Design 

- 2.2 Preliminary 

Design 

Acquisition1 

Construction and 

Deployment 

3.0 

- 2.3 Detailed 

Engineering I Design 

System Operations 

and Support 

4.0 - 

-3.3 Evaporator1 

Condenser 

Subsystem 

- 3.4 1 65-kWeTurbine 

Subsystem 

- 3.1 Engineering 

Design Control 

-4.1 System 

Operations 

- 3.2 Structural and -4.2 System Tests Facilities Support 

Systems 


The task leader is responsible for assigning all task activities and communicating task 

status and issues to the systems integration team through the systems .integration task 

leader. 

1.4 Exwriment Constraints 

The baseline HMTEAINPPE design is defined by technical, budget, and other factors. A 

major constraint is the capability of the seawater systems at NELH to deliver cold and 

warm water to the experiment. Because of a cooperative agreement between DOE and 

the State of Hawaii, resulting from their cost shares in constructing e resent seawater 

system at the NELH, federal project have priority in using 0.61 m /s 9600 gal/min) of 3 r warm surface seawater and 0.41 m /s (6500 gallmin) of deep-ocean cold seawater. 

Water temperature is in the range of 24.5OC to 28OC for the surface water (depending on 

the time of the year) and remains at 6.1°C for the cold water. Environmental permits 

have also been obtained for disposal of the seawater at these stated volume flows on an 

experimental basis. 

The ground level at the proposed experiment site is 4 m (13 ft) above mean sea level. 

Existing onshore piping at NELH supplying the seawater to the HMTEAINPPE site is 

0.61m (2 ft) in diameter and is projected to cause more hydraulic losses than are 

acceptable for assuring net power production. Thus the piping must be enlarged. 

Early projections of power consumption within the system indicated that up to 75% of 

the generated power will be used for the system operation with the earlier constraints. 

The small fraction of gross power available makes the technical objective of demonstrating net power production more demanding than in the case of a larger plant, where the 

net power fraction may approach 70% (Parsons, Bharathan, and Althof 1985). 

Scaling the experiment from 1.2 MWt for the HMTSTA to 7.0 MWt for the HMTEAINPPE 

represents a significant extension in the technology data base. In addition to the limited 

water available to the DOE project, the budget, schedule, and other external constraints 

argue in favor of the current scale--called the 165-kWe (7.0-MWt) experiment--for the 

HMTEAINPPE project. 

Given the combination of constraints, analytical studies of various experimental options 

have supported the selection of the baseline experimental system summarized in this 

report (Shelpuk 1985; Link and Parsons 1986; Lewis, Trimble, and Bowers 1987; PICHTR 

1988). 
CHAPTER 2.0 

SYSTEM INTEGRATION 

This chapter describes the status of the project at the end of the conceptual design. The 

design effort was carried out from January through October 1989. 

2.1 Goals and Objectives 

The NPPE will be developed in two steps: the HMTEA and the NPPE. To assure compatibility, it is important to define the goals and objectives in terms of both of these steps. 

The first step will be to construct the HMTEA. 

The goal in developing the HMTEA is to design, build, and operate an experimental 

system capable of performing OC-OTEC evaporator and condenser tests at a scale required for the NPPE. The purpose of these tests will be limited to verifying the design 

performance and assuring the reliable operation of the heat exchangers. Another goal is 

to have the HMTEA function as a development facility for full-scale evaporator and 

condenser components after the primary goal for the NPPE is attained. 

The goal of the NPPE is to experimentally verify the feasibility of OC-OTEC through 

validating power-generation techniques by producing at least 40 kW of net power with a 

minimum risk of failure. The intermediate completion of the HMRA directly supports 

the minimum risk requirements for achieving the NPPE goal. 

To meet the NPPE system goal, the following objectives were established (OET System 

Team Meeting Minutes 1989): 

Obtain data on subsystem and system performance using seawater in the HMTSTA/ 

HMTEAINPPE. 

Develop and validate necessary predictive codes. 

Obtain system control data and operational experience. 

Evaluate dynamic stability and transient responses of the system. 

Begin considering scaling issues. 

2.2 Design Requirements 

The design of the HMTEAINPPE is established by combining requirements of the project 

goals and objectives and the resource water availability at the test site at the NELH. 

The baseline system design conditions are in Table 2-1. 

The design requirements for the HMTEAINPPE are then described in the following areas: 

function, reliability, operation and maintenance, safety, cost, and aesthetics. 

The function of the NPPE is to produce net power at the stated level to evaluate the 

feasibility of the open-cycle concept. Analysis has shown that a minimum 200 kWe of 

gross power must be generated to meet the project goal of 40 kWe net power availability. Another functional requirement is that the system be fully controllable with 

built-in stability. The design should also provide for adequate performance margins so 

the system will operate satisfactorily during the off-design conditions expected during 

most of the testing period. 

The generated electrical energy will be dissipated in controlled resistive loads or delivered to the utility grid. A functional requirement important to the designer is that the 
Table 2-1. HMTEAINPPE Seawater System Design Conditions 

as of September 14, 1989 (units revised 12-5-89) 

Nominal High Low 

WARM SEAWATER: 

Flow rate 

Temperature 

Density 

Barometric head 

COLD SEAWATER: 

Flow rate kg/s 420 460 

gPm 6,500 7,150 

Offshore CW pipe total kg/s 860 860 

flow rate g Pm 13,300 13,300 

Temperature "C 

Density 

lb/f t 

Barometric head m 

ft 

CW pipe density head loss m 

ft 

MIXED DISCHARGE: 

Flow rate 

Temperature 

Density 

OTHER: 

Tide 

Barometric pressure 

m 0 t 0.61 

ft 0 + 2.0 

mbar 1,015 1,026 

psi 14.72 14.88 

2 2 Acceleration of gravity at 20 deg latitude: 9.786 m/s (32.1 1 ft/s ) 

Grade elevation at the NPPE: 4 m (13 ft) 
latter be done safely and within specifications and tolerances acceptable to the local 

utility company, especially when the NPPE is operating as a test bed for new and innovative components. 

Finally, although the net power criterion for the experiment will be applied in a powerproducing configuration, the system will be designed so 10% of the design steam flow can 

be condensed in a surface condenser to produce desalinated water. 

Reliability considerations are driven by a stated requirement that the system have a 

useful life of two years of intermittent operation, with a planned continuous operation 

for up to six months. Although no specific availability has been stated as a requirement, 

certain guidelines on the reliability were put forth at the first systems team meeting. 

No redundancies for major equipment are required; however, all components are to be 

chosen within the state of the art, with considerations given to reliability and field 

maintainability. One of the purposes of the planned long-term NPPE tests will be to 

evaluate the reliability and availability of the system. 

The system should be designed to require a minimum of down time for either scheduled 

or unscheduled maintenance, with no more than two persons needed for routine operation 

and maintenance. 

Safety is the highest priority item, with adequate design margins applied throughout, 

Control redundancy with built-in safety sensors and controls and annunciators will be 

required. A detailed safe operating procedure will be developed, maintained, and followed throughout the functional period of the HMTEAINPPE. Accompanying detailed 

analyses of the failure modes and operators' actions for the system will be developed 

during the design process. 

The cost of the project must be compatible with anticipated funding levels, which can 

vary but are now estimated at $15M over the three-year development period. 

Finally, aesthetic design will be used throughout, with emphasis on simplicity, accessibility, and interchangeability. Attractive use of simple but functional components and 

subsystems will be emphasized during design. 

2.3 System Analysis 

2.3.1 Analysis Method 

System analysis identifies the component operating conditions and specifications (constrained within an overall physical concept for the system) that will satisfy the system's 

technical and functional requirements at the minimum first cost and operating cost. For 

the NPPE a higher priority has been placed on minimizing the risk associated with 

producing net power than on achieving the minimum cost. The cost must be kept within 

the budget constraints, however. 

Validating system feasibility and analysis codes is also a prime objective of the experiments to be carried out with the HMTEAJNPPE. The NPPE tests will provide data to 

upgrade the design codes for use in optimizing the systems for a minimum cost in commercial installations. 

The design was initiated by completing several system analyses, which were used to 

eliminate unreasonable designs based on the criterion of maximum net power from a 

fixed quantity and quality of cold water. Early system analyses identified the sensitivity 
Table 2-2. Power Budget Distribution and Design Projections 

Task Description of Power Budget Design 

Power Use Goal Projections 

We) (kWe) 

A. Turbine/structures Generator output 

gross (minimum) 

less oil system 

B. Seawater systems Supply/discharge 8 1 

pumping (maximum) 

C. Heat exchangers Parasitic water 

pumping (maximum) 

D. Ancillary systems Vacuum, control, 

and intercoolers 

(maximum) 

System integration NETPOWER=A-(B+C+D) 40 

of system performance to the various design and operating parameters and established 

the boundaries of the experimental test envelope used in the HMTSTA tests (Zangrando 

et al. 1989). The results of the HMTSTA experiments were then used to establish nominal NPPE component specifications. These design performance estimates incorporated 

into a first-cut system definition are bracketed by experimental test results for the 

direct-contact heat exchangers, projected performance of the low-pressure turbines 

based on design studies (Mechanical Technology Inc. 1989; Fong, Masutani, and Neil1 

1989), and documented performance in the case of other commercially available equipment. 

The subsystem performance goals for the NPPE were developed based on the operating 

conditions that can be expected in a typical OC-OTEC plant. These goals (shown in the 

third column of Table 2-2) guided the first cycle of the design. An original assumption of 

cold-seawater temperature of 7OC was used in developing this table. Operational experiences with the upgraded seawater system at the NELH have resulted in updating the 

available cold-seawater temperature to 6.1°C. A detailed systems analysis effort described in Section 2.3.2 provides the next-level update of the subsystem performance 

projections (see Table 2-2, fourth column). 

2.3.2 Analysis Results 

A system performance analysis was then completed to establish the preferred steam flow 

through the turbine (Figure 2-1) and water flow through the flash evaporator 

(Figure 2-2). Figure 2-1 plots the projected gross power vs. steam flow through the 

turbine. Because gross power is proportional to the product of the steam flow and its 

temperature drop across the turbine, this figure shows that there is a specific steam flow 

rate that maximizes the power. Losses in warm-seawater pumping are less than in cold, 

so the temperature distribution in the power-maximizing system shifts toward a lesser 
Figure 2-1. Variation of gross power vs. steam flow for the net 

power-producing experiment (NPPE) 

temperature difference across the evaporator than in the condenser; i.e., more warm 

water is used than cold. The chosen value of 3.5 kg/s for the steam flow will accommodate reduced resource-water temperature differences (caused by seasonal variations 

in the ocean temperature), andlor less-than-expected heat-exchanger effectiveness with 

minimum decrease in the power output. The design value also will handle increased 

resource-water temperature differences or improved heat-exchanger effectiveness, or 

both, which would call for a higher value of steam flow through the turbine for maximum 

power. 

The exact value of turbine inlet and outlet temperatures is determined by heatexchanger design and performance. The results of the HMTSTA experiments on the 

direct-contact heat exchangers have guided the selection of the design performance 

goals for the system. Confirmation of the design in the planned HMTEA experiments is 

important to the success of the NPPE. 

Figure 2-2 shows the effect of water flow rate variation on system net power output. 

Increasing cold-water flow rate increases the net power and vice versa. However, the 

available cold-water flow rate is fixed for this project. To achieve maximum net power, 

all of the available cold water must be used. 

3 Figure 2-2 is also used to identify a warm-water flow rate of 0.61 m /s (9600 gallmin) as 

a preferred design value. A set of assumptions with respect to the onshore seawater 

supply piping has been made (for details see Chapter 4.0). Careful design of the onshore 

piping is necessary to preserve the performance gains identified in this analysis. 

Table 2-2 summarizes the nominal design conditions for the NPPE facility as established 

by the assumptions and the foregoing analysis. This table also compares the design 

projections against the initial goals set for the various tasks. The design results are the 

basis for the component specifications and hardware designs presented in the following 
chapters. Based on the results of these 

analyses, the original design goals can be 

compared with the present design projections. At this stage of the design, the 

gross power projection increased to 213 

kWe, and the net power increased to 84 

kW Most of these increases stem from 

up&titing the cold-water temperature from 

7°C to 6.1°C. 

Potentials exist for an increase of up to 

40 kWe in the system parasitic losses because of uncertainties in the performances 

of the seawater system and vacuum exhaust system. To assure a net power of 40 

kWe, the power output of the turbine/ 

generator unit must be maintained at over 

200 kWe. 

Table 2-3 provides a detailed set of performance projections for the various 

major components and subsystems of the 

NPPE (updated November 15, 1989). 

Cold 

I seawater I 

Warm 

I seawater I 1 

560 620 680 

Seawater flow rate (kgls) 

Figure 2-2. Variations of HMTEA/NPPE 

system net power with resource-seawater 

flow rates 

2.3.3 Sensitivity Analysis 

A sensitivity analysis was completed to establish the influence of the expected bands of 

uncertainty associated with the key variables in the design. Table 2-4 provides a 

summary of the results. The major variables are categorized broadly as the component 

performance, resource variability, and control variables. 

Shown first are the sensitivities to uncertainties in the projected heat-exchanger performance. For the evaporator, the design effectiveness is 0.90. Variations in its effectiveness of +0.05 and -0.03 are realistic and yield changes in the net power of +3 kWe and 

-4 kWe, respectively. The performance variations in the direct-contact condenser result 

in the net power variation of +3 kWe to -6 kWe. Other uncertainties in the components 

permit the determination of an average net power of 82 kWe, with a standard deviation 

of +5 kW . At this stage of the design, based on a statistical approach (Kline and 

~c~lintocz 1953), the average net power of 75 kWe is assured at 67% confidence level, 

and a minimum net power of 67 kW, is assured at a 99% confidence level at the design 

resource-water temperatures and flow availability. Validated experimental data on the 

seawater supply systems and performance projections of the vacuum exhaust system will 

affect these projections up or down as the design progresses. 

A major uncertainty currently is the projected performance level of the disposal trench. 3 The projections call for discharge of 1.02 m /s (16,100 gal/min) of spent water at a 

discharge pool elevation of 0.9 m (3 ft) above sea level. An elevation of 3 m (10 ft) to 

discharge this water will reduce the net power by 30 kWe, as shown in the second category 
Table 2-3. Nominal Design Conditions for the NPPE System (Projections at Nov. 1989) 

Component Description 

Turbine 

Generated power 

Generator efficiency 

Turbine efficiency 

Steam inlet temperature 

Steam outlet temperature 

Steam inlet velocity 

Steam outlet velocity 

Steam enthalpy drop (total to total) 

Steam flow rate 

Maximum tip speed 

Rotor outside diameter 

Rotational speed 

Diffuser 

Recoverv factor 

Outlet steam velocity 

Steam outlet temperature 

Evaporator 

Mist-removal pressure loss coefficient 

Mist-removal steam velocity 

Mist-removal pressure drop 

Evaporator-mist pressure loss coefficient 

Steam generation velocity 

Evaporator-mist pressure drop 

Liquid loading 

Steam generation temperature 

Evaporator effectiveness 

Warm-seawater inlet temperature 

Warm-seawater outlet temperature 

Warm-seawater flow rate 

Fraction of equilibrium gas release 

Noncondensable gas evolution 

Air leakage rate 

Spout-seawater velocity 

Evaporator platform area 

Mist removal area 

Direct-Contact Condenser 

Dif fuserlcondenser pressure loss coefficient 

Diffuserlcondenser pressure loss 

Cocurrent pressure loss 

Countercurrent pressure loss 

Cold-seawater inlet temperature 

Cold-seawater outlet temperature 

Steamlgas outlet temperature 

Outlet steam flow rate 

Cocurrent gas loading 

Cold-seawater flow rate 

Equilibrium gas release 

Noncondensable gas evolution 

Air leakage rate 

Water loading 
Table 2-3. Nominal Design Conditions for the NPPE System (Concluded) 

Component Description 

Condenser Exhaust Compressor Train 

Number of compression stages 

Intercooler/precooler pressure drop 

Compressor stage pressure ratio 

Intercooler gas exit temperature 

Motor/compressor efficiency 

Power requirement 

Warm-Seawater Flow System 

Onshore supply pipe, nominal diameter 

Supply pipe head loss 

Overall pumping efficiency 

Supply pump power 

Evaporator static loss 

Evaporator dynamic loss 

Discharge pipe head loss 

Discharge trench head loss 

Total warm-water head loss 

Supply pump head 

Overall pumping efficiency 

Supply pump power 

Discharge pump head 

Overall pumping efficiency 

Discharge pump power 

Total warm-water pumping power 

Cold-!Seawater Flow System 

Onshore supply pipe, nominal diameter 

Total supply head loss 

Density head loss 

Condenser static loss 

Condenser dynamic loss 

Discharge pipe head loss 

Discharge trench head loss 

Total cold-water head loss 

Supply pump head 

Overall pumping efficiency 

Supply pump power 

Discharge pump head 

Overall pumping efficiency 

Discharge pump power 

Total cold-water pumping power 

Miscellaneous Equipment 

Lube oil pumps 

Oil cooling system 

Plant control system 

Intercooler water pumps 

System 

Total parasitic power requirement 

Net power production 

'We 

m 

m 

m 

m 

m 

m 

m 

m 

m 

% 

k We 

m 

% 

'We 

k We 

'We 

k We 
Table 2-4. NPPE System Net Power Sensitivities 

Rotor 

Gross Total Change 

Projected Shaft Generator Parasitic Net in Net 

Design Uncertainty Power Output* Power Power Power 

Value Range (kWe) (kWe) (kWe) (kWe) (kWd 

Design 

Component Performance 

Evaporator effectiveness 

Condenser effectiveness 

Turbine (T-T) efficiency 

Exhaust system efficiency 

Generator efficiency 

Warm-seawater pumping (kWe) 

Cold-seawater pumping (kWe) 

Miscellaneous parasitics (k We) 

Seawater Disposal 

Trench elevation (m) 

Resource Variability 

Warm-water temperature. ("C) 

Control Variables 

Warm-water flow (kgls) 

Cold-water flow (kg/s) 

- 

"Generator output is (rotor gross power minus bearing loss) times generator efficiency. 
A third category for the sensitivity analyses is the resource variability. The influence of 

a variation in the warm-seawater temperature is projected to be the greatest, causing a 

net power change of -33 kWe at a low temperature of 24.5OC to +36 kWe at a high temperature of 27.5OC. Tides are not expected to influence the net power, because the level 

of cold seawater in the supply system rises and falls with the tide. Another variable is 

the cold-seawater temperature; however, this temperature is not projected to vary 

significantly as compared with that of the warm seawater. 

A fourth category shown in Table 2-4 is the resource-water flow rates, which can be used 

to control the system operation within a limited range. A range of 10% (increase and 

decrease) in the warm and cold water flow rates yields roughly a 1-kWe and 4-kWe (increase and decrease) change in net power, respectively. 

Demand for cold seawater from other users (such as the Hawaii Ocean Science and 

Technology [HOST] park) at the NELH will influence the parasitic 1 sses associated with 3 the offshore section of the water intake pipe. When only 0.41 m /s (6300 gal/min) of 

cold s water is drawn through the offshore pipe (as opposed to the design draw 'of 7 0.61 m Is), the net power yield from the NPPE system will increase by roughly 10 kWe 

above the design value, because the supply pump efficiency does not deteriorate at low 

heads. 

The variability in the warm-water temperature previously considered covers a range 

encountered in field measurements at the NELH over the past several years. This range 

of variability falls within the plateau of system performance described in Figure 2-1 and 

indicates a variation of steam flow rate from 3.2 to 3.7 kg/s of steam. Projected variation in the turbine efficiency over this range of steam flow is minimal. It is thus reasonable to expect a minimum of 40 kWe out of the plant at any time of the year. 

The projected variations in the design efficiency for the turbine are based on an extensive data base on which the turbine design is founded. It is expected that the stated 

turbine design efficiency goal will readily be met or surpassed. Thus, the technical 

objective of achieving a net power production of 40 kWe remains well in reach. 

2.3.4 Code Benchmark Comparisons 

Two types of mathematical codes were used for system analysis. A detailed code 

entitled OC-OTEC-SYS was developed at SERI and extensively used at SERI and at 

PICHTR. This code accepts the basic operating parameters for the system as inputs 

along with limited component design assumptions. This code exercises all of the design 

variables to arrive at complete system and component designs that maximize power 

output or minimize system cost, depending on the chosen criterion. 

Early heat-exchanger and turbine development studies have shown that these components 

can be designed for high performance over a wide range of operating conditions. It is not 

necessary to complete a design analysis of each component every time a system analysis 

effort is undertaken. Rather, the key components can be described as black boxes in the 

system analysis (using a given efficiency) with resource-water flow rates and temperatures as the input parameters. The output defines the system operating conditions, 

intercomponent flow rates, temperatures, and power. This simple second model can also 

be installed in a spreadsheet format; the results of varying the key parameters are then 

obtained quickly, with preferred operating conditions arrived at by observing trends. 

The results reported in this section were obtained using the simple model. The results 

were then compared with those obtained from OC-OTEC-SYS on several benchmark runs, 
one of which is reported in Table 2-5 (projections as of September 18, 1989). The correlation between the two methods is sufficiently close that the results are indistinguishable. 

The chosen design for the NPPE conservatively defines components for the critical 

systems. Any design margin caused by this approach will provide for unanticipated 

performance shortfalls caused by factors not modeled. 

The first cycle of the design was completed by projecting the performance of critical 

components based on experimental data. The performance requirements then permitted 

the sizing and assembly of those and other components in a manner to maximize net 

power. 

2.3.5 System Control 

Control of an OC-OTEC system has not been analyzed in detail; however, it appears not 

to be a significant issue. There are four control requirements: long-term control to 

accommodate seasonal temperature variations, component performance changes, and 

tidal influences; short-term control to handle sudden load or water flow variations resulting from HOST park operations; normal system start/stop sequences; and emergency 

shutdown. 

Long-Term Control. Variations that occur over periods of days or weeks are readily 

accommodated by this system. It may not be necessary to provide active controls because the expected fluctuations will not drive the operation of all components and the 

system far from their design point. Resource-water flow rates will be regulated by 

variable-speed drives on the supply system. Discharge water pumps will be controlled 

using water discharge pool levels in the heat exchangers to maintain those levels within 

specified limits. Such a system will be self-adjusting and deliver more or less power, 

depending on the nature of the variation. Further analysis will be carried out during the 

preliminary design stages to assure stable system operation. 

Short-Term Control. Short-term control presents a different type of problem. Load 

variations are likely; power output fluctuations from the generator should be accommodated by corresponding variations to the load bank using the electrical switching circuitry. Sudden load losses or gains are typical requirements for power plants of all types, 

and standard switchgear is well able to handle the requirement. 

Sudden fluctuations of water flow, however, would cause a different type of situation. 

Because the NPPE will share the water flow from the NELH cold-water system with the 

HOST park facility, fluctuating demands from HOST users would set up water flow 

oscillations in the experiment. Preliminary assessments indicate that these oscillations 

occur over time periods of 5 to 10 minutes and are well damped at startup or gradual 

shutdown. Some overshoots in the pump-sump pool depression would occur. Head variations in the supply sumps may lead to possible unstable pump operation, pump cavitation, 

and potential trapping of air bubbles in the offshore pipe. The implications of the shortterm sump oscillations will have to be assessed during the preliminary design. 

Wave action may not affect the supply of cold water but may influence the supply of 

warm water. The high flow resistance through the warm-water supply system in comparison with the wave excursions should mitigate the potential problems. 

The short-term fluctuations in the flow of warm or cold water should not affect the 

rotational speed of the turbine rotor because of its large inertia; however, further 
Table 2-5. Systems Code Benchmark Comparison of NPPE Design Conditions 

Projections by Systems 

SERI PICHTR Analysis 

(91 18/89) (9113189) Basis 

Turbine and Generator 

Generated power Model with radial 

turb., 9/ 13/89 

H. Link 8/89 

D. Bharathan 

5/8/89 

Computer model 

9/13/89 

Computer model 

9/ 13/89 

Computer model 

9/13/89 

Computer model 

9/ 13/89 

Generator efficiency 

Turbine efficiency 

Steam inlet temperature 

Steam outlet temperature 

Steam inlet velocity 

Steam outlet velocity 

Steam enthalpy drop 

(total to total) 

Steam flow rate Computer model 

9/13/89 

275 m/s 

2.9 m 

1800 rpm 

Maximum tip speed 

Rotor outside diameter 

Rotational speed H. Link 8/89 

Diffuser 

Pressure recover factor Y (2(p2 - P~)/P~V~ 

D. Bharathan 

5/8/89 

Eff = 0.87 

Computer model 

9/13/89 

Computer model 

9/13/89 

Computer model 

9/ 13/89 

Computer model 

9/ 13/89 

Inlet steam velocity 

Outlet steam velocity 

Steam inlet temperature 

Steam outlet temperature 

Evaporator 

Mist removal pressurf loss 

coefficient (2bplpv ) 

Mist removal steam velocity 

TRC /I3 

4/ 1 2-41 13/89 

Computer model 

9/ 13/89 

Computer model 

9/13/89 

Mist removal pressure drop 

Evaporator/mist elimina r pressure 5' loss coefficient (2bplpv ) 

Steam generation velocity D. Bharathan 

5/8/89 
Table 2-5. Systems Code Benchmark Comparison of NPPE Design Conditions (Cont'd) 

-- 

Projections by Systems 

SERI PICHTR Analysis 

(91 18/89) (91 13/89) Basis 

Evaporator (cont Id) 

Evaporator/mist eliminator 

pressure loss 

Liquid loading F. Zangrando 

8/14/89 

Computer model 

9/13/89 

F. Zangrando 

8/14/89 

OEP 3/3/89 

Computer model 

9/13/89 

OEP 3/3/89 

TRC 113 

4112-4/13/89 

Computer model 

9/13/89 

NPPE 

2122-2/23/89 

F. Zangrando 

8/14/89 

Computer model 

91 13/89 

Computer model 

91 13/89 

Steam generation temperature 

Evaporator effectiveness 

Warm-seawater inlet temperature 

Warm-seawater outlet temperature 

Warm-seawater flow rate 

Fraction of equilibrium gas release 

Noncondensable gas evolution 

Air leakage rate 

Spout-seawater velocity 

Evaporator platform area 

Mist removal area 

Direct-Contact Condenser 

Dif f user-condenser pressure loss coef f. 

Diffuser-condenser pressure loss 

L. Neil1 8/89 

Computer model 

9/ 13/89 

H. Link 8/89 

H. Link 8/89 

OEP 3/3/89 

Computer model 

911 3/89 

Computer model 

9/13/89 

Computer model 

9/13/89 

H. Link 8/89 

OEP 3/3/89 

1 /ef f??? 

H. Link 8/89 

Computer model 

9/13/89 

H. Link 8/89 

Computer model 

9/ 13/89 

Cocurrent pressure loss 

Countercurrent pressure loss 

Cold-seawater inlet temperature 

Cold-seawater outlet temperature 

Steam/gas outlet temperature 

Outlet steam flow rate 

Cocurrent gas loading 

Cold-seawater flow rate 

Actual/theoretical flow rate 

Fraction of equilibrium gas release 

Noncondensable gas evolution 

Air leakage rate 

Water loading 
Table 2-5. Systems Code Benchmark Comparison of NPPE Design Conditions (Cont'd) 

Projections by Systems 

SERI PICHTR Analysis 

(91 18/89) (911 3/89) Basis 

Condenser/Exhaust Compressor Train 

Number of compressor stages 8. Parsons 

8/ L 1/89 

B. Parsons 

811 1/89 

Computer model 

9/ L 3/89 

Not part of model 

Intercooler/precooler pressure drop 

Compressor stage pressure ratio 

Intercooler gas exit temperature 

Motor/compressor efficiency 8. Parsons 

8/L 1/89 

8. Nichols 

not modeled 

Power requirement 

Warm-Seawater Flow System 

Onshore supply pipe, nominal 

diameter 

Supply pipe head loss 

H.J. Green 

8/22/89, 36" steel 

H.J. Green 

8/22/89 

vertical wet pit 

H.J. Green, 

stat. + dynam. 

incl. above 

H.J. Green 

8/22/89 sep. 

disch. 

H.J. Green 

7/13/89 

Computer model 

9/ 13/89 

Evaporator static loss 

Evaporator dynamic loss 

Discharge pipe head loss 

Discharge trench head loss 

Total warm-water head loss 

Supply pump head 

Overall pumping efficiency, 

supply pump Supply pump power 

Discharge pump head 

Overall pumping efficiency, 

discharge pump 

Discharge pump power 

Total warm-water pumping power 

H.J. Green 

7/13/89 

H.J. Green 

7/13/89 

Computer model 

9/13/89 

Cold-Seawater Flow System 

Onshore supply pipe, nominal 

diameter 

0.61 m 0.61 H.J. Green 

8/22/89 

24" HDP DR26 

4.6 m 4.17 H.J. Green 

8/22/89 

vertical wet pit 

Total supply head loss 
Table 2-5. Systems Code Benchmark Comparison of NPPE Design Conditions (Concluded) 

Projections by Systems 

SERI PICHTR Analysis 

(91 18/89) (9113189) Basis 

Cold-Seawater Flow System (cont'd) 

Density head loss Computer model 

9/13/89 

H.J. Green 

7/13/89, stat. + 

dynam. 

incl. above 

H.J. Green 

8/22/89 

Sep. disch. 

H.J. Green 

7/13/89 

Computer model 

9/13/89 

Condenser static loss 

Condenser dynamic loss 

Discharge pipe head loss 

Discharge trench head loss 

Total cold-water head loss 

Supply pump head 

Overall pumping efficiency, 

supply Pump Supply pump power 

Discharge pump head 

Overall pumping efficiency, 

discharge pump 

Total cold-water pumping power 

H.J. Green 

7/13/89 

H.J. Green 

7/13/89 

Computer model 

9/ 13/89 

Miscellaneous Equiprnent/Essential Aux. 

Lube oil cooling system 

Oil cooling system 

Plant control system 

Intercooler water pumps 

R. Davis 9/89 

R. Davis 9/89 

R. Davis 9/89 

R. Davis 9/89 

System 

Total parasitic power requirement Model + projections 9/13/89 

Model + projections 9/13/89 

Net power production 

analysis must be carried out to determine the effects on the turbine efficiency caused by 

changes in steam velocity from such fluctuations. 

Short-term control requires a detailed analysis to ascertain the control requirements and 

means for implementation. Proper management of the total NELH water delivery system is required. Special attention should be paid to a quantitative analysis of the shortterm control aspects during preliminary design. 
Start/Stop Control. The startup and shutdown of the HMTEA will follow procedures 

similar to those used at the HMTSTA (Parsons et al. 19891, and no specific additional 

control requirements are foreseen. For the NPPE, however, start-up and shutdown involve synchronizing the turbinelgenerator. Details of the control requirements are 

discussed in Chapter 3.0. 

Emergency Control. The NPPE can be shut down in an emergency by venting (by breaking the vacuum) the working fluid space to atmospheric pressure at a controlled rate. No 

detailed calculations have been made on the spin-down time of the turbine rotor nor the 

aerodynamic heating that will occur as the turbine alters function from an expander to a 

compressor. However, given the kinetic energy of the rotor at the operating speed, 

preliminary calculations indicate that spin-down will occur in 10 to 20 minutes. Spraycooling may be needed to prevent significant temperature rise and damage to the internal system from frictional heating. A positive stop valve is not planned in the steam 

supply line because of the large size required and the flow losses it might introduce. A 

complete safety analysis, during normal operation and shutdown and in emergency situations, will be necessary to determine the extent of shutdown controls required. 

The controls and safety issues were considered in a qualitative fashion during conceptual 

design and do not appear to be of any real concern. A priority activity for preliminary 

design will be to complete a quantitative analysis of these issues and to initiate an analysis of major transients that might affect either the performance or reliability of the 

NPPE components. 

2.4 Physical Integration 

2.4.1 Integration Options 

The first iteration of the design assumed no component interface losses. With the low 

steam pressure used in this cycle and the large volume of the process steam and water 

flows, careful attention is necessary to the physical arrangement of the components and 

definition of the flow path. Two basic configurations for the overall system were considered. The first involves a horizontal orientation for the turbine axis of rotation with 

the evaporator and condenser located on either side of the turbine. The second option is 

defined with a vertical-axis turbine with the condenser located below and directly on the 

axis of the turbine. The evaporator is located in an axisymmetric annular cylindrical 

space with its inside diameter in line with the turbine inlet perimeter and its outer 

diameter extending to the outer wall of the vacuum enclosure. Figure 2-3 provides 

illustrations of the two options. 

2.4.2 Selection Criteria and Rationale 

The design choices of all of the HMTEAINPPE components are based on their performance, technical risk, schedule, and cost in the given order of importance. The following 

discusses the relative merits of the choices in the physical integration of the system. 

Performance. System analysis has shown that the proof of the feasibility (for producing 

net power at the stated power levels possible with the resource seawater flow available 

at NELH) requires near-peak performance from all subsystems used in the experiment. 

The operating margin for achieving the net power goal of 40 kWe is, at best, 25% of the 

gross power under an optimistic scenario and zero with conservative projections of 

component performances. The requirement for minimal performance losses is a key 

consideration for the physical system arrangement. 
Horizontal-axis turbine, 

"telephone" design 

cn 

- 50 

I 40 a, 

> 

m 

0 loft Grade 

10 

60 Vertical-axis turbine, 

"dome" design 

Axisymmetric arrangement 

Half-sectional view 

cn 

H 40 a, 

> 

0 

a 

m 

a 2 

m 

Technical Risk. Falling short of the 

predicted performance goal is the 

major risk factor and is also considered in the performance category in 

assessment; so the technical risk 

carries double weight as used. Under 

this category, the possibility is considered that the experiment will not 

work at all or will fail prematurely. 

Experimental research completed to 

date provides high confidence that 

the system will operate as projected 

in either configuration. Therefore, 

consideration of the technical risk is 

confined to evaluating the possibility 

of system or component failure or 

degradation over the useful life of 

the experiment. 

5 2.1 ' nofi MsL 

Schedule. Consistent with the research agenda outlined in the OET 

10 Grade Multiyear Program Plan (U.S. DOE 

1985), the NPPE project should be 

completed within the time frame of 

the schedule as presented in Figure 

1-3, with an acceleration of that 

Figure 2-3. Sketches of lltelephonevl and "dome" schedule being desirable. Any slipconfigurations for HMTEAINPPE page of schedule will also affect 

overall system cost. 

Cost. Developmental cost for the experiment should be consistent with the projected 

federal and PICHTR budgets, a requirement which appears feasible at this point. Any 

cost increase from the baseline projection could put the project at risk. The ability to 

raise additional funds, however, is perceived possible; a potential increase in cost would 

be less risky to the project than would its failure to achieve either the stated performance or reliable operation. Project cost details are discussed in Section 2.5. 

2.4.3 Configuration Options 

The two options considered for physically integrating the HMTEAINPPE are described 

below: 

Telephone. The telephone configuration is based on an extension of the HMTEA design 

studied by an industrial design team (Dillingham 1987b). The concept is based on expanding the existing HMTSTA design to a larger size and mounting it on a fabricated steel 

tower. To accommodate a turbinelgenerator for an NPPE, the side-by-side heat exchangers need to be separated horizontally to locate the turbine between them. The 

ducting space must be adequate to assure proper steam flow distribution across the face 

of the turbine and to allow for low-loss steam exit from the evaporator and entrance to 
Figure 2-4. General arrangement of components in "telephone" 

configuration for HMTEA/NPPE 

the condenser. These considerations result in a structure schematically shown in 

Figure 2-4. The advantages of this approach are as follows: 

The heat exchangers reflect the construction and configurations previously built and 

tested. 

Vacuum containment made of steel should be reliable and easy to maintain for extended periods. 

The arrangement provides flexibility for any potential future modifications. 

The disadvantages are as follows: 

The structure is large, requiring a minimum footprint area of 14 x 5 m (42 x 15 ft). 

To provide for adequate seismic protection, the turbine foundation needs to be extended directly to bedrock. 

No overhung desi n (bearings located only on one end) exists of a single-rotor turbine 

of about 5000 kg 7 11,000 lb) with a horizontal axis of rotation; such a design may cause 
a significant imbalance in bearing load and is a departure from standard engineering 

practice. 

A horizontal-axis single-rotor design that is overhung represents a departure from all 

prior open-cycle designs, which use a balanced double-rotor horizontal-axis design or a 

single-rotor vertical-axis design. 

To ensure uniform flow distribution to and from the turbine, two large volutes to guide 

the steam flow are needed. Each volute is estimated to wei h 10,000 kg (22,000 Ib). 

Evaporator and condenser vessels are roughly 4.6 m (15 ft 7 in diameter, approaching 

the diameter of the dome. The evaporator poses difficulties in arranging the multiple 

spouts; the condenser requires a separate countercurrent vessel(s). 

All vessel walls and steam piping must withstand full internal vacuum, requiring stiffened steel wall construction. 

Dome. This concept is based on a vertical-axis turbine and a concrete vacuum enclosure 

proposed and studied in earlier open-cycle OTEC design efforts (Westinghouse 1979; 

Block and Valenzuela 1985; Valenzuela et al. 1988; Penney et al. 1984; Ludke and Chow 

1984). In this concept, the system is built around the vertical axis of the turbine with all 

of the major components contained within a reinforced-concrete cylindrical enclosure 

that also acts as the mounting pedestal for the turbine. (Designs by Valenzuela et al. and 

Ludke and Chow use a double-rotor horizontal-axis turbine enclosed in the dome structure.) The concept is defined in the artist's sketch shown in Figure 2-5. 

The advantages of this concept are . - 

It is a compact arrangement, sharing several structural functions in a single structure 

and simplifying the on-site assembly construction. 

The straight in-line exhaust of steam from the turbine to the condenser minimizes flow 

losses, provides excellent steam distribution to the condenser, and facilitates the 

integration of a diffuser. 

Heavy rotors of large sizes (such as in hydraulic machines) are typically arranged with 

a vertical axis to minimize radial bearing-load imbalances. 

The vertical shaft allows preloads on the radial bearings to be adjusted in the field to 

achieve the most effective rotor dynamic performance. 

Diffuser and other steam piping needs to withstand only the pressure corresponding to 

that between the evaporator and the condenser; they can be made of lightweight 

plastic materials and can readily be modified. 

The water distribution manifolds can be integrated readily within the vacuum structure 

using plastic plumbing. 

The disadvantages are 

The vacuum containment, if made of concrete, requires special coatings to prevent air 

permeation and seawater leaching. 

Access to internal parts in a larger system may require heavy equipment and complicated procedures. 

Based on a qualitative comparative evaluation of the two options at the first working 

group meeting (SERI 1989), the systems team selected the dome configuration as the 

prime configuration for the NPPE system with both steel and concrete as candidate 

materials. Further design work was directed to proceed with the dome configuration. 

The telephone configuration was to be maintained as a backup option in the event that 

future work reveals serious technical or cost issues related to the dome's further 

development. 
2.4.4 The Integrated NPPE 

System Design 

System Description. The conceptual 

design of the NPPE is defined in the 

overall system drawing shown in Figure 

2-6, the process flow diagram in Figure 

2-7, and the system specification 

summarized in Table 2-6. A 

photograph of a scaled model of the 

vacuum structure and internals is 

shown in Figure 2-8. 

The system is built around a verticalaxis, mixed-flow turbine rotor 2.9 m 

(9.5 ft) in diameter supported by a 

7.3-m (24-ft) outside-diameter vacuum 

housing that also acts as the structural 

backbone of the system. An exhaust 

diffuser is incorporated into the Figure 2-5. General arrangement of 

turbine design to recover most of the components in "dome" configuration for 

kinetic energy of the discharge. This HMTEAJNPPE 

diffuser, with a conical center body, is 

2 m (6.6 ft) in diameter at the turbine 

exhaust and 3.65 m (12 ft) long and forms an annular entry to the cocurrent condenser 

section that is 4.1 m (13.5 ft) outside diameter and 2.4 m (8 ft) inside diameter. The 

diffuser is contained within the vacuum enclosure and, thus, is required to withstand the 

minimal pressure drop across the turbine and to support its own weight. It can be 

fabricated from light-gauge steel or possibly a plastic composite. 

The electrical generator is located above the turbine assembly and outside the vacuum 

enclosure. It is an air-cooled, 4-pole, 1800-rpm synchronous generator with a brushless 

exciter. Further details on the turbinelgenerator are found in Chapter 3.0. 

The turbine is supplied with low-pressure steam from an arrangement of 12 vertical 

spouts, each 0.24 m (10 in.) in diameter and located so the spout exit is 4.6 m (15 ft) 

above grade level. The spout evaporator occupies an annular cylindrical space bounded 

by the vacuum enclosure and an annular inner wall that separates the evaporator and the 

condenser. Warm seawater is brought to the evaporator by a single pipe 0.76 m (30 in.) in 

diameter that penetrates the vacuum enclosure at the bottom and rises in the vacuum 

space to an evaporator water manifold located just below the spouts. A chevron-type 

mist eliminator is located above the evaporator to assure that moisture impingement is 

not a problem for the turbine rotor. 

The condenser is direct contact with coaxial cocurrent and countercurrent stages fabricated and filled with a cooling tower packing (Munters Plasdek CF-25060 fill or equivalent). The first stage of the condenser occupies the outer annular cylindrical space 

connected to the diffuser exit and operates in a cocurrent flow of steam and water 

vertically down. Cold seawater is supplied to the condenser packing from a series of 

radial water manifolds equally spaced around the perimeter of the condenser space. A 

cold-water supply pipe 0.6 m (24 in.) in diameter penetrates the vacuum enclosure at the 

bottom and delivers the water for both condenser stages. The distribution manifold 

located on the centerline of the system is sized to distribute appropriate flows to the two 

stages. The second stage of the condenser is centrally located within the first stage. It 


Table 2-6. HMTEAINPPE System Specifications 

TurbineIGenerator 

Radial inflow, single stage 

Steam consumption 

Steam inlet temperature 

Steam exhaust pressure 

Turbine speed 

Rated power 

Generator design output 

Voltage/phase/cycles 

Shaft seal leakage 

Heat Exchangers 

Evaporator 

Type 

Seawater flow 

Seawater temperature 

Steam generation rate 

Steam temperature 

Design effectiveness 

Direct Contact Condenser 

Type 

Seawater flow 

Seawater temperature 

Steam consumption rate 

Steam inlet temperature 

Inert gas content in steam 

Design water effectiveness 

Surf ace Condenser 

Type 

Seawater flow 

Seawater temperature 

Steam consumption rate 

Steam inlet temperature 

Inert gas content in steam 

Design water effectiveness 

Seawater Pumps 

Function 

1.38 kPa 

1800 rpm 

300 kWe 

220 kWe 

480VI3ph AC/60 Hz 

5 0.0005 kg/s 

Vertical Spout 

Packed Column 

420 kg/s 

6.1°C 

Plate-Fin 

42 kg/s 

6.1°C 

Warm- Cold- Warm- ColdWater Water Water Water 

supply supply Discharge Discharge 

Number 1 1 1 1 

Flow rate (kg/s) 620 420 620 420 

Total head (m) 1.4 3.6 3.2 5.4 

Efficiency (%) 6 6 7 2 7 2 7 2 

VoltageJphase < ------------ 480~/34ph ........................ > 
Table 2-6. HMTEA/NPPE System Specifications (Concluded) 

Ancillary Systems 

Vacuum Exhaust Compressors 

Type: Centrifugal and/or positive-displacement compressors, 

4 stages, yith precooler and interstage coolers 

Inlet flow rate (m 1s) 6.9 

Inlet pressure (kPa) 1.2 

Discharge pressure (kPa) 10 1 

Maximum power (kWe) 40 

Instrumentation TBD 

Control System Specifications TBD 

Vacuum Structure 

Size 

Material 

Leak rate 

Internal pressure 

External loading 

Design requirement 

6.4 m ID x 7.6 m height 

Coated reinforced concrete 

0.0005 kg/s 

1 kPa (absolute) 

40 tons (vertical) 

Seismic zone 3 

operates in countercurrent flow to condense additional leftover steam in the exhaust so 

the pumping requirement for the vacuum vent system is minimized. The countercurrent 

stage is constructed of the same material as the cocurrent stage but is longer with an 

overall packing height of 0.91 m (3 ft). 

A surface condenser is used in the system to condense up to 10% of the steam flow to 

produce desalinated water as a by-product of power generation. The surface condenser 

will be developed as a test article to acquire performance data. The condenser is fabricated as a brazed aluminum structure, with extruded channels forming the seawater 

passages and offset finstock acting as the enhanced condensing surface in the vapor 

stream (Panchal and Genens 1989). This surface condenser is located outside but next to 

the vacuum enclosure. It is supplied with cold seawater tapped from the main condenser 

coolant supply and uses separate water circulation pumps on the desalinated water and 

seawater streams. Chapter 5.0 provides further details on the heat exchangers. 

The seawater systems consist of supply, discharge, and distribution portions for the warm 

and cold waters. The distribution system to the evaporator and condenser was described 

earlier. The major part of the supply system already exists at NELH. This existing 

supply system will be modified by adding onshore piping and replacing existing pumps 

with a complement of high-efficiency pumps. One supply pump will be used in each of 

the warm and cold water streams. The supply pumps will be located at the intake sump 

I_, and will be of a vertical wet-pit construction using mixed-flow impellers. The housings 

will be coated cast iron with yet-to-be-chosen impeller and shaft materials. Another 

pair of pumps will provide for the discharge and will be barrel-type pumps located near 

but outside the vacuum enclosure. All pumps will use variable-frequency drives for flow 

control. 
Figure 2-8. Photograph of vacuum structure showing internal 

arrangement of HMTEA/NPPE system components 
Warm-water onshore piping will be 0.91 m (36 in.) in diameter; cold-water onshore piping 

will be 0.61 m (24 in.) in diameter. The material for these pipes is unspecified at this 

stage. On the discharge side, the water streams will be combined downstream of the 

pumps into a 1.22-m- (48-in.)-diameter polyethylene pipe, which will be buried and will 

end in a percolation discharge trench. A proposed test trench on the site is 3.05 x 3.0 x 3 30.5 m (10 x 10 x 100 ft) in its excavated dimensions to discharge nominally 0.19 m 1s 

(3000 gallmin.). This trench is expected to be operational in early 1990. The size of the 

trench needed for the HMTEAINPPE discharge will not be determined until construction 

and testing are completed on the performance of the test trench. An existing water 

disposal sump on the site will not be used. For further details on the pump designs and 

other remaining open issues, see Chapter 4.0. 

The most important ancillary equipment in the HMTEAINPPE is the vacuum compression 

system. The design established for the HMTEAINPPE uses a high-efficiency, multiplestage, intercooled compressor. The system will use a precooler and direct-contact or 

surface-type interstage coolers with four compression stages, operating with nearly a 3:l 

compression ratio and projected 60% design efficiency in each stage. Potential suppliers 

for the vacuum compressors were identified during the conceptual design. Resolution on 

the type of compressors to be used was achieved. However, assessment of off-the-shelf 

machines suggested that to maintain low power consumption, the compressor(s) will be an 

engineered product and will use positive-displacement and/or centrifugal rnachine(s). A 

high-priority effort during the preliminary design is warranted to finalize the vacuum 

system. More details on the vacuum system and other ancillary systems are provided in 

Chapter 6.0. 

For increased safety, the physical locations of the critical high-speed rotating machines, 

namely the turbine and vacuum pumps, will be higher than 6.4 m (21 ft) above grade level 

to ensure that no seawater damage could occur during any unanticipated operating conditions. 

I 

Process Description. Surface seawater enters the evaporator held at an absolute pressure of about 2.6 kPa, where it generates low-pressure steam by flash evaporation. 

Nearly 0.6% of the incoming water is converted to steam. The spent water is drained 

from a pool below. Dissolved gases, mostly oxygen and nitrogen, also evolve from the 

seawater. Almost 100% of these gases are projected to be released from the seawater. 

The mixture of these gases and steam flows upward, carrying some seawater droplets 

generated during evaporation. 

A mist eliminator located directly above the evaporator allows water droplets of over 

40-micrometer diameter to be removed from the steam. Steam then enters the stator of 

the turbine where it is accelerated to match its exit tangential velocity to the tip speed 

of the rotor. Thus, the steam enters the rotor radially and exits the rotor vertically 

downward. The power imparted by the steam to the rotor is transmitted to the generator 

via a flexible coupling. 

A diffuser located below the rotor allows partial recovery of the kinetic energy of the 

spent steam. Below the diffuser, the steam contacts cold seawater distributed over 

packings in the cocurrent section of the condenser. Nearly 80% of the steam is condensed here, causing the water temperature to rise about 5OC. The remaining steam is 

collected centrally and flows upward through a countercurrent section of the condenser. 

This section is also provided with cold seawater from above, distributed over packings. 

Up to 98% of the remaining steam is condensed in this section. The spent water from 

both sections of the condenser is collected in a pool below and discharged from the 

vacuum vessel. 
Dissolved gases are also released from the cold seawater. The remainder of the steam, 

together with the noncondensable gases (released from both seawater streams and leaked 

into the vessel), is removed through a central pipe connected to the vacuum exhaust 

compressors located outside but adjacent to the vacuum vessel. 

The power generated is distributed to water pumps, vacuum pumps, other ancillary 

equipment, and essential control elements. Excess power from the system will be dissipated in a load bank or fed into the local utility grid. 

2.4.5 Site Plan 

The site plan for the HMTEAINPPE involves three primary locations. The pump station 

for the existing seawater systems is located approximately 70 m (200 ft) from the shoreline and currently contains 12 pumps in two sumps. The low-head system that will be 

used for the HMTEAINPPE contains three pumps to supply the cold-water flow and four 

pumps to provide the warm-water flow in a separate sump. As was discussed earlier, this 

arrangement will be reviewed in the later design phases to achieve higher pump efficiency and reduced piping losses. Although the final decision has not been made, the new 

supply pumps will be located in or near the present sumps. 

The main experiment location is between the NELH administration building and the site 

of the HMTSTA, as shown on the plot plan in Figure 2-9. The domed cylindrical apparatus and a HMTEAINPPE control building will be installed on a pad at this location. The 

seawater supply piping from the pump station to the experiment is 200 m (650 ft) long 

and will be buried. 

The piping run from the experiment to the third HMTEAINPPE site, namely the discharge trench, is 1.22 m (48 in.) in diameter and 215 m (700 ft) long. The discharge pipe 

will also be buried until it emerges in the trench. The trench site is north of the other 

facilities at the NELH; the tests at the test trench will allow determination of the ultimate trench size. 

2.4.6 System Interfaces 

A vital function of the systems integration task is to establish and manage the interfaces 

among all the components and subassemblies that make up the experiment. These interfaces include size and location of physical connections, flow rates, electrical power, 

instrumentation, and utility hookups. Key interfaces among the various tasks as outlined 

in the work breakdown structure are identified in Table 2-7 and are noted in the accompanying process flow diagram shown in Figure 2-7. This list is by no means complete. 

Early in the preliminary design, closure points on all of these interfaces are compatible 

and realistic. Once a closure occurs, changes in interfaces will only be possible through a 

prescribed formal procedure. Such a procedure will assure the various designers that 

unknown or arbitrary changes in their interface requirements will not occur. As the 

design proceeds, requirements may arise on both sides of an interface that are incompatible or that compromise safety, reliability, accessibility, or maintainability. In those 

cases, the system integrator will define, document, and control the alterations in the 

interfaces. 

2.5 Cost Projections 

The objective of a conceptual design is to develop data such that a skilled reviewer can 

assess the project's technical merit and potential risks and can roughly estimate the cost 

of turning the concept into hardware. At the present level of the HMTEAINPPE design, 

Table 2-7. List of Critical System Interfaces 

Qualitative 

From* To* Number* Medium* Description 

Quantitative 

Description 

Warm seawater 

supply to 

evaporator 

Warm seawater 

drain from 

the evaporator 

Cold seawater 

suppiy to 

condenser 

Cold seawater 

drain from 

condenser 

!'Mist-free" steam 

supply to turbine 

Spent steam from 

turbine to firststage condenser 

"Mist-free1' ? 

low-pressure steam 

supply to surface 

condenser 

Three-phase 

electrical output 

from generator 

Vacuum-rated flange, 30" 

dia., to handle up to 

10,600 gallmin 

Vacuum-rated flange, 36" 

dia., to handle up to 

10,600 gal/min 

Vacuum-rated flange, 24" 

dia., to handle up to 

7,150 gal/min 

Vacuum-rated flange, 30" 

dia., to handle up to 

7,150 gallmin 

Vacuum-rated horizontal 

flange, 21' dia., to 

handle up to 3.9 kg/s 

steam flow at a pressure 

of nominally 2.4 kPa. 

Physically located 25' 

above grade 

Vacuum-rated horizontal 

double flange, 8' 

ID* on inner; 13.5' 

OD* on outer, to 

handle up to 3.9 kg/s 

steam at nominal pressure 

1.4 kPa. Physically 

located 13' above grade 

Vacuum rated vertical 

flange, 24" dia., to handle 

0.35 kg/s steam flow at 

pressure 1.4 kPa. 

Physically located at 

21' elevation 

Rating 300 kWe, 480 V, 

3 phase, 60 Hz AC 
Table 2-7. List of Critical System Interfaces (Concluded) 

Qualitative Quantitative 

From* To* Number* Medium* Description Description 

H A 9 LS Exhaust from 

condenser 

Cold-water supply 

to surface 

condenser 

Cold-water drain 

from surface 

condenser 

H A 12 LS Exhaust from 

surf ace condenser 

Vacuum-rated horizontal 

flange 24" dja., to handle 

up to 3.8 m /s of lowpressure steam inert 

mixture 

Vacuum-rated flange, 

10" dia., to handle up to 

750 gallmin cold seawater 

Vacuum-rated flange, 

10" dia., to handle 

up to 750 gal/min cold 

seawater 

Vacuum-rated vertical 

flange to handle steam 

and noncondensable 

flow, 8" diameter 

*See below for explanation of abbreviations 

I = System integration (000) 

H = Heat exchanger ( 100) 

T = Turbine (200) 

S = Seawater system (300) 

A = Ancillary system (400) 

WW= warm seawater LS = low-pressure steam 

CW= cold seawater DW = desalinated water 

CS = control signal E3 = 3 phase, 480 V, electrical supply 

IN = instrumentation signal El = 1 phase, 120 V, electrical supply 

CA = compressed air OD = outer diameter 

LO = lube oil ID = inner diameter 

a number of alternatives remain to be decided upon for several components. Details of 

possible interrelational aspects among the various components have not been thoroughly 

considered. Possible cost premiums associated with site-related special requirements, 

and the details of the civil structures needed, have not been considered. This lack of 

detail makes a comprehensive cost estimate impractical. 

A detailed analysis of the design and cost of the major components in the NPPE has been 

completed and is documented in this report. Some costs were established by parametric 

relationships or analogy. Because the costs of major components (which form the bulk of 

the system cost) were evaluated closely, even though others are estimated roughly or 

overlooked altogether, the total cost is adequately defined for budgetary planning. 
The following cost estimate is broken down into two categories: the cost of design and 

construction, and the extent of project management support necessary to complete the 

intermediate research experiments and analysis and to establish the management basis 

for constructing and operating the experiment. 

2.5.1 Design and Construction 

Table 2-8 on pages 44 and 45 summarizes the estimated cost of completing the design 

and construction of the NPPE. The costs are organized according to the major 

component and subsystem categories discussed in this report. Each category is broken 

down into three cost elements: equipment purchases; contracts for design, installation, 

and repair or modification; and consultants to provide expert assistance in specifying and 

evaluating the HMTEAJNPPE and its elements as well as analyzing and troubleshooting 

the equipment and experimental results. As is apparent in this report, the design and 

cost are most thoroughly defined for the turbine, ancillary systems, seawater systems, 

and heat exchangers, Because these elements account for more than 75% of the total 

cost, a contingency ii;nd of approximately 15% of the total cost is applied that could 

cover an uncertainty of up to 100% in the cost of the structure and the projected 

extension to the disposal trench (the major uncertainties at this point). 

2.5.2 Technical Management 

Because the HMTEAJNPPE is an ongoing research experiment with intermediate experiments to be completed with the HMTEA, "technical management'' may be a misnomer for 

the supporting technical effort that must be provided. Some of the activities that must 

be completed by the design team for the NPPE to be successful are as follows: 

Studies and Analyses. At the end of conceptual design, numerous choices need to be 

made regarding component specifications, materials, operating parameters, and so on. 

As the design and interfaces begin to solidify, conflict and compromises that arise can 

only be resolved through objective analysis. Such studies and analyses will have to be 

undertaken as required on a continuing basis. 

Subcontract Management. The technical, coordinating, and administrative functions for 

each subcontract will involve one or more team members. 

Configuration Management and Interface Control. The chief engineer will formalize and 

manage the documentation of the design and assume that all interfaces are compatible. 

Project Management and Review. Overall coordination and control of budget and schedule will have to be maintained. Timely reports to DOE will be required on the progress 

of the project and any corrective or preventive actions needed. 

Test Plan Development. A comprehensive HMTEA test plan will be developed and test 

results will be evaluated before the NPPE test plan is finalized. However, an early draft 

of the NPPE test plan will be made available to guide the design team on the NPPE 

testing requirements. 

Other HMTEA/NPPE Plans and Documentation. The following plans and documentation 

will be developed and managed in the course of the project: 

a Work breakdown structure 

a Reporting procedures for cost and schedule 
Procurement plan, including inspection, quality control procedures, and acceptance 

procedures and criteria 

Permits and licensing plan, including environmental monitoring and reporting 

Safe procedures and operating plan that will include analyses of component failures 

and recommended actions under unusual or emergency situations 

Construction methods and equipment plan that will assure that the HMTEAJNPPE will 

meet standards of quality, aesthetics, and safety appropriate at the NELH 

Operation and maintenance plan that assures that procedures necessary for safe operation and maintenance of the useful life of the system will be followed 

Equipment disposition plan for continued use or disposition of HMTEAJNPPE hardware 

and associated systems and equipment. 

Other activities that may arise in the course of the project development and that are 

appropriate for the OTEC project team will be completed within the technical management function. 

It is estimated that a full-time force of 12 workers, distributed among the three cooperating organizations, will be needed to carry out these various activities. This estimate 

contains a contingency of 50%. 

2.6 Remaining Issues and Future Work 

2.6.1 Technical Issues 

The objective of the system analysis and integration task is to provide the analysis, 

engineering, planning, and direction to address the following specific issues of the 

HMTEAINPPE: 

System performance 

Component interfaces 

System control 

System stability and dynamics 

Hardware design and integration. 

All of these issues were considered in varying levels of detail during the conceptual 

design. System stability and control were examined mostly qualitatively. Definition and 

analysis of the control schemes remain as outstanding issues at this stage. 

The compatibility of component interfaces was examined, and all interfaces appear to be 

satisfactory for good performance and reliable operation. The beginning of preliminary 

design is the appropriate time to clearly define and document all of the interfaces and to 

begin to formalize the procedures for their management. 

Performance of Developmental Items. The structural design of the vacuum containment 

vessel needs further definition, particularly regarding air leakage and impact of seismic 

activity. The vessel serves the dual function of vacuum containment and structural 

support of the turbine. Adequacy of the structure for both of these requirements will 

have to be evaluated in the light of a possible seismic event. 

The planned tests to evaluate the permeability of the trench that is under development 

will provide data on the extension required to discharge the entire HMTEAJNPPE water 

stream and on the associated parasitic power penalties for the system. Should the permeability prove to be unacceptable, alternative methods for disposing of the discharge 

streams need to evaluated and ranked and assessments made of potential environmental 

effects and costs. 


The preliminary design and development of the vacuum exhaust system must be monitored closely to ensure its acceptability for parasitic power consumption and timely 

delivery at or below the projected cost. 

The design of the advanced plate-fin surface condenser for producing desalinated water 

must be evaluated. Close attention should be paid to the compatibility of the material 

with seawater, the potential for corrosion and pitting, and the projected life expectancy. 

The effects of these developmental items on performance, technical risk, cost, and 

schedule of the system must be monitored especially carefully throughout their developmental period. 

Other Critical Areas. Other areas of concern lie in the projected off-design performance 

of the power generating system, the potential effects of critical component failures on 

safety of personnel and integrity of the system, the effect of seismic action and 

tsunamis, and other environmental causes and issues. All of these areas require detailed 

evaluation during the next stage of the HMTEAINPPE design. 

2.6.2 Preliminary Design 

The next step in the project is to proceed to preliminary design and to put in place the 

described technical management activities. 

The objective of preliminary design is to obtain adequate data with supporting drawings 

and analyses to establish the engineering and economic feasibility of the concept and to 

accommodate all technical uncertainties through alternative engineering approaches. 

The design will be sufficiently detailed that cost and schedules can be established to 

within 25%. 

To achieve the objective, the following activities and documentation are required: 

A drawing of the preliminary site plan that shows the experiment's location at the 

NELH, establishes all utility interface requirements, defines all future expansions (if 

required), and shows access to workshops, offices, storage, and sanitary and other 

facilities that may be required 

A final drawing of the overall layout of the system showing size and location of all 

major subsystems drawn to scale and showing major dimensions 

A preliminary piping and instrumentation drawing providing an engineering estimate of 

the amount, type, and cost of the$-umbing, electrical, and instrumentation and control 

components 

a Final system performance specifications defined as major subassemblies. These specifications will establish operating parameters (e.g., temperatures, flow rates, pressures, 

and levels), minimum performance measures, and outline drawings for each major 

subassembly. 

An analysis of the system performance at the design point and over a range of expected operating conditions. This analysis will use validated or consensus models and 

methods with all critical performance assumptions traceable to manufacturers' data or 

valid experimental research results. 

An analysis of safety, expected life, reliability, and maintenance requirements for the 

system operating continuously or intermittently at NELH over a period of two years 

An analysis of system control, performance, stability, and data-acquisition requirements, with special attention paid to off-design and emergency conditions 

An itemized construction cost estimate accurate to within 25% for the system as 

installed and ready to run 
A preliminary test plan that meets the basic objectives of the project and allows for 

additional testing that can provide data for understanding the system's operation and 

designing future systems if they are perceived as being valuable. 

2.6.3 Final Design 

The design cycle will be completed with the final engineering design and specification for 

the HMTEA/NPPE. 

The objective of this phase is to produce design and construction drawings and specifications so the schedule and cost for procuring, building, installing, and operating the 

HMTEA/NPPE hardware can be quoted on a firm fixed-price basis. 

The design and documentation package that will permit procuring, constructing, and 

successfully operating the experiment will include 

A final approved drawing package providing adequate detail that every hardware item 

in the system may be purchased or built. The package will include all drawings necessary for site preparation, utility interfaces, component and subsystem subassembly and 

installation, and assembled system checkout. 

Final specifications and data sheets for all system components and instrumentation 

A detailed safe operating procedure 

A fixed-price cost estimate and construction bid package 

A system acceptance procedure that defines criteria by which the DOE and its costsharing partners will accept ownership of the system. The procedure will specify the 

responsibilities of all parties and make provisions for resolving any disagreements that 

may arise. 

A final test plan that will identify data, instruments, test procedures, and data analysis 

methods suggested for the experiment 

A detailed technical report that addresses the analyses used as the basis for the design, 

with particular emphasis on areas that needed supplementation or modification from 

the preliminary design results 

A final data package reflecting the as-built key system parameters. 

It is recommended that the final design package be reviewed by a competent architectural and engineering firm with established expertise in developing systems of similar 

complexity before the sponsors formally approve it. When the final design and approvals 

are completed, construction and installation can begin. 
CHAPTER 3.0 

TURBINE AND STRUCTURES 

3.1 Development of the Turbine 

3.1.1 Goal and Objective 

The objective of the NPPE turbine task is to develop a turbinelgenerator unit that will 

ensure net power production with minimum technical risk. 

The performance goal is generation of at least 165 kWe. Given improved operating conditions when 6.1°C cold seawater is used (earlier estimates were for 7°C cold seawater) 

and projected high thermodynamic efficiency for the turbine configuration chosen, the 

turbinelgenerator should be capable of exceeding its performance goal by as much as 

40 kWe. 

3.1.2 Approach 

The approach for this task (shown in Figure 3-1) is as follows: 

8 Evaluate the potential of existing turbine hardware for meeting the task goal at a 

lower cost or in less procurement time than a new design would require. 

8 Evaluate the expected performance of a turbinelgenerator unit using existing technology but new hardware. 

8 Choose the most appropriate turbinelgenerator option (existing hardware or existing 

technology) for procurement. 

8 Complete a detailed design of the turbinelgenerator, ensuring that it is physically and 

operationally integrated with the rest of the NPPE system. 

Fabricate the turbinelgenerator and perform factory tests so that required on-site 

assembly and checkout activities are minimal. 

8 Ship the unit to the test site, erect it on the NPPE structure, and make final 

adjustments. 

Validate performance with comprehensive acceptance tests. 

To ensure timely completion of this work, a turbine design team was formed. At the 

time of writing, members are H. Link, SERI; K. Fong, PICHTR; T. Rabas, ANL; and 

C. Castellano, U.S. DOE. Others who have contributed to the effort include T, Penney, 

B. Shelpuk, and D. Bharathan of SERI and S. Masutani of PICHTR. 

Consultants from the turbine industry have also played key roles in writing the specifications and statements of work for the Phase I and Phase I1 design contracts. Helpful evaluations of Phase I designs of existing hardware and technology were given by 

William Steltz, William Sanders of Turbo Technic Services, Inc., and Neville Rieger of 

Stress Technology Incorporated. Dr. Taher Schobeiri of Texas A&M University reviewed 

thermodynamic performance of radial-inflow and axial machines for this application. 

3.1.3 Options Considered and Eliminated 

3.1.3.1 Existing Hardware vs. Existing Technology 

The primary consideration was whether to procure a turbine that uses existing components such as blades or to build a new one based on new knowledge about design. A 

parallel-path development plan was enacted to assist in evaluating these two options. 
Phase 1 

Competitive design 

Turbine C bids d (MTI) Phase II Phase Ill Phase Ill 

requirements Fly-off - design - fabrication/, acceptance 

I D'd SOW installation test . Market + Phase 1 

survey design 

(MHIIPICHTR) BA-GO540801 

Figure 3-1. Approach to turbine development 

SERI staff members and a team of turbine consultants prepared a Request for Proposal in 

which commercial firms were asked to bid on a three-phase turbine development program. Approaches for both existing hardware and existing technology were available to 

the bidders. Two proposals were received for the existing-technology approach, and a 

subcontract was awarded to Mechanical Technology Inc. (MTI), for Phase I design. MTI 

proposed the use of a radial-inflow turbine--a configuration with which the company has 

had considerable experience. 

PICHTR investigated the existing-hardware option more fully so that a tradeoff evaluation could be made after Phase I design was complete. After investigating a number of 

options, PICHTR determined that Mitsubishi Heavy Industries (MHI) would best meet the 

the NPPE turbine requirements with existing hardware. MHI recommended use of blades 

from an axial-flow machine which had been designed for geothermal applications. 

Details of the two designs are contained in two Phase I design reports (MTI 1989; Fong, 

Masutani, and Neil1 1989), which were reviewed by the turbine task team, the turbine 

consultants, and T. Schobeiri (1989). The turbine team and consultants met at SERI in 

July 1989 for a final evaluation and selection. Results of the evaluation are summarized 

in Table 3- 1. 

The probability for net power was the deciding factor in the evaluation. The turbine consultants judged that MHI's efficiency prediction was overly optimistic, considering that 

the blades would be subjected to high incidence angles and negative reaction at the root. 

Design conditions for the MHI blades deviate significantly from standard designs, in 

which incidence is normally less than 10' and reaction is not permitted below +15%. MHI 

based its predictions regarding efficiency on existing performance data but, for proprietary reasons, would not present parts of that data for review. 

The MTI design was judged to have a good probability of achieving total-to-total isentropic efficiencies of 85% to 90%, which would result in a power output of 183 to 195 

kWe under steam conditions used in the Phase I design. 

3.1.3.2 Horizontal vs. Vertical Axis 

A second consideration was whether the turbine should have a vertical or horizontal axis 

orientation. This issue is not limited solely to turbines; axis orientation significantly 

affects structural design and could also influence design of the heat exchangers. 
Table 3-1. Preliminary Design Evaluation for Turbine 

Existing Existing 

Technology Hardware 

Design team MTI MHIIPICHTR 

Configuration Radial-inf low Axial 

Projected cost 

(Phase I1 and 

Phase 111) 

Delivery 36 months 19 months 

Assessment of probability 

of net power High 

Recommended for Phase I1 

design X 

Low 

A decision was made at the first working group meeting to proceed with the design of a 

"dome" structure rather than the tvtelephonevv structure. More discussion of the implications of this decision is provided in Chapter 2.0 of this report. Briefly, its effect on 

turbine design was to increase the desirability of a vertical-axis orientation, which integrates well with a dome structure. 

Structural integration issues aside, MTI and others identified other advantages of the 

vertical-axis turbine. First, the journal bearings do not have to support the weight of the 

turbine wheel and thus can be adjusted to meet rotordynamic requirements. Second, both 

journal bearings can be located on one end of the wheel in an "overhung" configuration 

without significantly increasing bearing loss. The overhung configuration eliminates the 

need for bearing support struts at the turbine exducer (outlet), which would impede 

steam flow and reduce efficiency of the diffuser. 

A disadvantage is that a larger thrust bearing is necessary to support the weight of the 

rotor and the aerodynamic thrust of the blades. Friction in this larger bearing consumes 

more energy than is required for the thrust bearing in a horizontal-axis machine. 

3.1.4 Other Major Considerations 

3.1.4.1 Radial-Inf low vs. Axial Configuration 

Both radial-inflow and axial configurations were considered for this turbine. Either 

configuration can be used to obtain high thermodynamic efficiency under the steam 

conditions expected for the NPPE. MTI, the winner of the turbine development contract, 

selected radial inflow for one or more of the following advantages: 

a Under the specified steam conditions, Reynolds numbers in the radial machine tend to 

be higher, reducing skin friction losses and slightly increasing efficiency. 

a Loading on radial blades is less, because the pressure drop is distributed over a larger 

area of the blade. 

a Blade-root connections are much larger and therefore not as susceptible to failure 

from fatigue or corrosion or both. 
3.1.4.2 Modified Steam Conditions 

Steam conditions for the turbine were modified twice from those first specified in the 

Request for Proposal, reflecting new information developed during conceptual design. 

The first modification (Table 3-2) increased the pressure drop across the turbine and 

slightly decreased steam flow. This change centered the nominal operating conditions 

within the range of uncertainty in performance of the heat exchanger. 

Table 3-2. Steam Inlet and Outlet Conditions for the NPPE Turbine 

Request for Phase I Phase I1 

Property Proposal Design Design 

Inlet Conditions 

Pressure (Pa) 

Saturation temp. ("C) 

Steam flow (kg/s) 

Moisture content (%) 

Air content (%) 

Enthalpy (kJ/kg) 

Outlet Conditions 

Pressure (Pa) 

Saturation temp. ("C) 

*Calculated for this table--not originally specified. 

**n/s: not specified. 

The second modification was made after the Phase I turbine design was completed. It 

increased both pressure drop and steam flow to reflect new estimates of the temperature 

of the cold-water resource. Earlier estimates of 7.0°C were based on NELH records of 

water supplied by the 0.30-m (12-in.) cold-water pipe. With higher flows through the new 

1 .O-m (40-in.) pipe, temperatures measured at the HMTSTA averaged about 6. I "C. Figure 3-2 shows gross power output from the generator as a function of steam flow rate. 

The new cold-water inlet temperature increases the power output at all steam flows and 

shifts the maximum power output, resulting in a slightly higher rate of steam flow. 

The turbine designer, MTI, has stated that the new steam conditions can be accommodated without resizing the turbine; however, detailed analysis may indicate a potential 

1% increase in turbine efficiency if the rotor wheel were slightly enlarged. 

3.1.4.3 Speed Control 

Speed control for steam turbines normally consists of both a stop valve and a bypass 

valve. In OC-OTEC, both valves would need to be quite large to be effective or to prevent excessive loss of pressure. Analysis of the turbine's transient response indicated 

that adequate speed control can be provided by controlling water flow rate, changing 

vent compressor flow, bleeding air into the evaporator or condenser, or changing electrical load. These indirect methods can be considered in lieu of steam control valves 

because the turbine's rotational inertia is very large compared to the driving force of the 

steam. MTI estimates that 30 seconds would elapse before a 20% change in load (or 
Steam flow (kgls) 

Basis: 

Figure 3-2. Operating conditions for system and turbine 

steam driving force) would cause the turbine to deviate 1% from its design speed. This 

very slow response is relatively easy to control, based on measurement of the phase angle 

between the generator electro-motive force (emf) and power grid voltage. 

3.1.4.4 Variable-Position Stator Vanes 

Hydraulic control of stator vanes is another means of speed control. In addition, it 

permits adjustment of the angle of the stator vane to maximize performance over a relatively wide range of steam conditions. This feature is attractive in an experimental system. However, the option was rejected because of added cost (about $500,000) and 

reduced performance caused by leakage around the stator blades. 

3.1.5 Design Performance and Uncertainties 

Total-to-total isenthalpic efficiency--a key performance parameter for the 

turbinelgenerator unit--describes the ratio of the work obtained from the turbine (not 

counting bearing and seal losses) to the amount that would be obtained if the turbine had 

no aerodynamic losses and all kinetic energy in the exhaust gas were recovered. 
MTI estimated that the turbine's total-to-total aerodynamic efficiency would be between 

85% and 90% based on the performance of similar turbines. Their loss analysis indicated 

85.3% to 86.8% total-to-total efficiency with a 7-m-long diffuser. They estimate other 

losses as follows: bearing and seal friction, 12.6 kWe; generator efficiency, 92%; and 

electrical parasitics for equipment such as the lubricating oil pump, 1 kWe. These 

estimates result in gross power output of 183 to 195 kWe and an overall (not totalto-total) efficiency of 74%. 

Since MTI1s Phase I design was completed, SERI has estimated that the turbine system 

would produce about 217 kWe, based on 

the revised steam conditions given in Table 3-2 

a total-to-total efficiency of 85% 

a a shorter (4-m-long) diffuser (preferred because it reduces 

overall height of the vacuum vessel) 

a bearing and seal losses equal to those estimated by MTI 

a 92% efficient generator. 

As noted above, MTI estimates 

that total-to-total efficiency 

could drop by 1% if the turbine is 

not resized for the new steam E951 rDesir 

conditions (see Figure 3-3). The 

shorter diffuser is likely to exhibit g .- a lower pressure-recovery factor .g 85 

(0.65 vs. MTI's 0.78), which would 

decrease total-to-total efficiency 360 - 

by 0.9%. The combined effect of - 

a these two changes would be to re- 5 

- 

duce estimated total-to-total ef- 320' 

ficiency to between 83% and f - 

88%. SERI's estimate falls in the 2 280, 

middle of this range. 3 - a 

3 - 0 

Bearing and seal losses should re- 240- 

main the same if the turbine shaft $ - 

is not resized. Although the gen- 200 - 

erator's efficiency is likely to increase as its capacity is 3.6 ' 

increased, no allowance for such 

an increase was included in the 3.4 - 

estimate. Electrical parasitics (U 

will be higher than estimated by 

MTI when effects are included of 1 .Turbine inlet constant at design 

cooling water pump power and condlion 

2.Bearing losses not included maintenance of buffer seal pres- 3.Efficiency evaporator flange to 

sure. SERI increased the estimate condenser flange, total-to-static 

of electrical parasitics from 1 to 2.8 - 

2.5 kWe. I I I I I 1 

1.6 1.7 1.8 1.9 2.0 2.1 

The uncertainty in SERI's esti- Pressure ratio 

mate of 217 kWe is on the order 

of +5%, which translates to a Figure 3-3. Turbine off-design performance 

gross power from the turbine/ characteristics 
generator unit ranging from 204 to 225 kWe. The draft specification written for Phase I1 

and Phase I11 of the turbine procurement contract requires an output of 215 kWe to 

provide a small margin below SERI's estimated performance. This specification may be 

revised during contract negotiations. 

3.1.6 Design Specifications 

The original technical specifications in the Request for Proposal for the turbine 

development contract were the basis for the full specifications for the Phase 11 turbine 

design (Appendix A of MTI 1989). MTI modified them to reflect improved knowledge of 

the turbine design that resulted from their Phase I design effort. The turbine task team 

further modified them to reflect new steam conditions and to identify performance 

requirements rather than physical configurations where appropriate. A final 

specification will be developed as part of the detailed design effort for use in procuring 

and testing the hardware. 

3.1.7 Layouts 

Figure 3-4 shows the main features of the turbinelgenerator unit. Standard commercially available components are used wherever possible. The rotor is made up of 26 

13-mm (0.5-in.) steel vanes welded to a steel hub. The hub is bolted to a 304-mm (12-in.) 

cantilevered vertical shaft. The shaft is supported by two sets of self-acting (i.e., hydrodynamic), oil-lubricated, tilting-pad journal bearings and a rolling-element thrust 

bearing. The shaft is sealed against air infiltration by a labyrinth seal and a double 

circumferential floating ring seal. 

A flexible-disk coupling between the turbine and the generator requires no maintenance 

and can be disassembled without removing the generator. The generator is a salientpole, synchronous machine that is directly driven by the turbine shaft. The generator can 

act as a motor when operated below synchronous speed to help speed up the turbine during start-up operations. 

The inlet plenum for the turbine occupies most of the dome of the vacuum vessel and 

transmits the turbine's weight to the outer walls of the structure. It has been configured 

from 25-mm (l-in.) steel plate with reinforcement ribs added at various positions. The 

fixed-blade assembly (stator) has 23 blades set to a nominal 20' angle, which can be 

adjusted during final checkout to maximize performance for the achieved steam flow 

conditions. 

The rotor casing and diffuser resist small loads caused by pressure differences between 

evaporator and condenser steam. The rotor casing will be constructed of steel plate and 

machined to provide a close clearance to the rotor vane. The diffuser can be constructed 

of light (6-mm) steel plate or composite plastic materials. 

3.1.8 Interfaces 

The turbine interfaces with the structure at three flange connections (Figure 3-5)--one 

between the turbine inlet plenum and the structure's outer wall, and two between the diffuser outlet and the direct-contact condenser. The turbine manufacturer will fabricate 

the lower flange on the outer wall to mate with the flange at the inlet plenum to ensure 

a leak-tight seal. 

The transition is abrupt between the diffuser outlet and the cocurrent stage of the 

direct-contact condenser because the diffuser is short, a desirable feature which 


minimizes the overall height of the vacuum structure. It has been estimated that the 

transition will cause very small pressure losses and will not significantly affect steam 

distribution in the condenser. In later design efforts, the turbine and heat exchanger 

teams will consider adding fairings to smooth the transition. 

Functional interfaces are shown in Figure 3-6. The sections entitled "Resistive Load 

Bank" and "System Switchgear and Protective Relays" were considered to be outside the 

scope of the original turbinelgenerator system. However, because of the numerous control and power interconnections required, the turbine team judged that risk could be 

reduced by combining these sections into the turbinelgenerator system. 

3.1.9 Schedule 

Timely procurement of the turbine is critical for completion of the NPPE on schedule. 

Various ways of accelerating the procurement schedule were considered, including negotiating the construction subcontract simultaneously with Phase I1 design. The approach 

that seems to combine the earliest delivery of the turbine with the lowest risk of major 

price increases is to combine Phase I1 design with construction, installation, and acceptance testing activities and to provide incentives for early completion. Table 3-3 shows 

the anticipated schedule at the time of writing. 

3.1.10 Cost and Uncertainties 

The Phase I design contract was $240K for the existing-technology option. The design 

team of MHI and PICHTR accomplished the Phase I design of the existing-equipment 

option for $40K. 

MTI estimated a cost of $2,40OK ?20% for a contract that combines Phase I1 design with 

construction, installation, and acceptance test activities. The design part of this amount 

is $600K. 

Other costs primarily support the consultant team. In-house labor for monitoring of the 

subcontracts and other activities is not included in the breakdown that is presented in 

Table 3-4. 

3.2 Structures 

3.2.1 Goal and Objective 

The objective of the structures task is to design and construct the vacuum vessel and 

support facilities for the HMTEAJNPPE. 

The vacuum vessel affects system performance by permitting air to leak into the vacuum 

system. This process adds to the noncondensable gases 'desorbed from seawater, 

increases vacuum compressor power, and decreases the system's net power. The goal of 

the design is that leakage into the vacuum structure add no more than 0.5 g/s of air to 

the vacuum system. If this goal is met, net power production will be reduced only 

0.5 kWe from a case in which the vessel is leak tight. 

Besides preventing air from entering the system, the vacuum vessel houses the directcontact heat exchangers and supports the turbine. It must provide for internal water 

supply manifolds and seawater discharge pools for the evaporator, the direct-contact 

condenser, the mist eliminator, and the turbine exhaust diffuser. 

Table 3-3. Schedule for NPPE Turbine 

Phase I design October 1, 1988 - August 1, 1989 

Phase I1 design 

Prepare Statement of Work August 1, 1989 - November 15, 1989 

Negotiate Phase I1 November 15, 1989 - January 1, 1990 

and Phase I11 

Engineering design January I, 1990 - October 1, 1990 

Phase I11 fabrication 

Purchase long-lead items August 1, 1990 - September 1, 1990 

Fabrication and assembly September 1, 1990 - June 1, 1991 

Factory tests June 1, 1991 - September 1, 1991 

Ship and install September 1, 1991 - January 1, 1992 

On-site tests January I, 1992 - June 1, 1992 

Turbine ready June 1, 1992 

Table 34. NPPE Turbine Costs (in $K) 

Task 

-- 

Equipment Contracts Consultation Total 

Phase I design 0 280 100 380 

Phase I1 design 0 600 5 0 650 

Fabricationltesting 1,600 0 100 1,700 

Turbine-related 

electrical items - 210 - - - 

Total 1,810 880 250 2,940 

The control building houses instrumentation and control functions, electrical power distribution cabinets, offices, and a visitors' center. 

3.2.2 Approach 

The approach for this task is as follows: 

Determine the applicability of Dillingham Construction's (1987b) existing HMTEA 

design (Figure 3-7) to current NPPE requirements. 

Choose the general configuration for the vacuum structure (dome or telephone) and 

develop specifications for diameter, height, internal features, and penetration. 

e Award a subcontract for design of the vacuum vessel. 

Determine requirements for the control building and other structures. 

Design the control building and other structures. 

e Determine whether needs for lifting equipment can be met using NELH equipment, 

renting local equipment, or purchasing new equipment. 

Award subcontracts to construct the vacuum vessel, control building, and other 

structures. 

A team of staff members from SERI, ANL, and PICHTR was formed to accomplish this 

work. This report summarizes their work during the conceptual design stage of 
Figure 3-7. A partial representation of the structures' development. Team 

Dillingham design of the HMTEA members at the time of this writing 

are H. Link, SERI; K. Fong, PICHTR; 

and L. Genens, ANL. Previous team 

members who contributed to this 

conceptual design are D. Bharathan 

of SERI and S. Masutani of PICHTR. 

3.2.3 Options Considered 

and Eliminated 

3.2.3.1 Dome vs. Telephone 

Configuration 

The most significant consideration 

was the choice between a dome or 

telephone configuration for the vacuum vessel. The telephone configuration features separate vessels for 

the evaporator and direct-contact 

condenser with the turbine located in 

a connecting pipe between the two. 

The present HMTSTA is an example 

of a telephone configuration. The recently designed HMTEA (Dillingham 

Construction 1987b) is a second 

example. SERI proposals made in 

1987 (Trenka 1987) for the HMTEA 

showed that the Dillingham design 

could be modified to operate as a 

NPPE. 

The dome configuration encloses the . .. evaporator and direct-contact 

condenser coaxially within one \ structure. Westinghouse ( 1979) designed a 100-MWe floating plant based on a dome configuration. This configuration was 

also suggested by Creare, Inc. (Valenzuela et al. 1988) for a 10-MW land-based plant and 

by T.Y. Lin and SERI (Penney et al. 1984) for a 1-MWe OC-O&C experiment. The 

systems team chose the dome configuration for continued development at the first 

working group meeting in February 1989. Additional analysis by the heat-exchanger 

team (and documented in Chapter 5.0 of this report) provided additional support for the 

dome configuration. 

Structurally, the primary advantage of the dome configuration is that the surface area 

for the vacuum barrier is minimized. Using the most recent numbers for heat-exchanger 

planform areas and heights, the dome configuration would use almost 30% less surface 

area than the telephone configuration. This minimizes cost and reduces potential for air 

leakage. A second advantage is the elimination of two external elbows (or involutes) and 

an external diffuser. These components are complexly shaped and would be costly to 

fabricate. They also require large flange connections to the heat-exchanger vessels, 

which could increase air leakage. 
The primary advantage of the telephone configuration is easier access to the evaporator 

or condenser to replace major components. With the dome, the turbine and its inlet 

plenum must be removed to install or remove major components inside the vessel. With 

the telephone configuration, only the overhead elbow would require removal. 

3.2.3.2 Materials for Vacuum Vessel Walls 

Three options were considered for the vacuum vessel walls: steel only, reinforced concrete only, and a combination of the two. Steel provides an excellent barrier to air leakage and is thus the most common material for small vacuum vessels. For larger sizes, 

the combination has been used; concrete provides the strength and steel the impermeability required for hi h-vacuum vessels. The largest known vacuum vessel in the United 5 States is the 4,000-m pentagonal Planetary Aeolian Laboratory at the Ames Research 

Center (a NASA research center in Ames, California), which features concrete-only 

walls. Any of the three options could be considered for the HMTEAJNPPE. 

In the NPPE, the vacuum vessel must support not only dead weight and vacuum loads, but 

also the large rotating mass of the turbine. It is critical, therefore, that the vessel not 

exhibit resonant vibration frequencies around the 30-Hz (1,800-rpm) rotational frequency 

of the turbine. Although steel-only walls withstand other structural loads with a plate 

thickness of 28 mm (1.125 in.), finite-element analysis indicated that the vessel would 

have a resonant frequency close to 30 Hz. To increase the stiffness sufficiently to 

ensure a resonant frequency of at least 110% of the turbine excitation, a plate thickness 

of 44 mm (1.75 in.) is required. ANL directed the design and analysis of this steel-only 

vessel (Figure 3-8). They estimated the cost to be $294,000. The high cost and difficult 

on-site assembly of such thick steel plates reduce the attractiveness of this option. With 

the low cost of reinforced-concrete construction, the vacuum vessel walls could be 

constructed as thick as 18 in. and still cost significantly less than steel walls. Estimates 

for cost of a concrete-walled vessel range from $120,000 to $180,000 (Oswald 1989; 

Bharathan memorandum 1989; Flug 1989). 

Although a finite-element analysis of concrete walls has not been completed, calculations indicate that 18-in. walls would have natural frequencies over 40 Hz--well past the 

frequency range of concern. In addition, the added mass and higher damping characteristics of concrete minimize vibration and noise at all excitation frequencies, so the vessel 

would seem more "solid." 

Regarding concrete, the two biggest concerns are ease of modification and susceptibility 

to increased leakage. Although steel would probably be easier and less expensive to modify than concrete, major changes are not anticipated for this apparatus. 

Penetrations have been configured and located to facilitate a variety of modifications 

without requiring new penetrations. For example, the access penetrations are configured 

to double as possible ventilation ports, observation ports, or feedthroughs for instrumentation wiring. Should a new large penetration be required, methods and equipment have 

been identified (Stallman 1989) to permit such a modification without major cost. 

Brackets and fixtures can be securely mounted on a concrete wall with anchor bolts. 

Such mountings do not present much more difficulty than the required cleaning, welding, 

and repainting associated with attachment of brackets to a steel wall. 

The second concern with concrete, rate of leakage, is affected by 

Permeability 

Leakage through construction joints and around penetration inserts 

Leakage through seals 

Leakage through small cracks. 

Studies (Westinghouse 1979; Valenzuela et al. 1988) have indicated that, although concrete is permeable, permeation through structural walls would be insignificant compared 

to the rate at which noncondensable gases evolve from seawater in an OC-OTEC plant. 

In addition, permeation can be reduced to negligible levels if inexpensive coatings are 

applied to the outside of the vessel. 

Experts from structural engineering companies (Stallman 1989; Stinnett 1989) have indicated that leakage around penetrations and through construction joints can be controlled 

with standard methods used in concrete water tanks, including elastomeric water stops 

and sealants. 

Seals are likely to be the biggest cause of air leakage. However, the choice of wall 

material probably will not affect leakage through seals, because it is expected that metal 

inserts will be used for penetrations and for the flange at the connection to the turbine. 

A significant concern is that cracks may form as a result of shrinkage or one of the seismic tremors that occur frequently on the big island of Hawaii. Depending upon the construction details, such a crack is not likely to result in leakage if it does not extend 

completely through the wall. If it does, the reinforcing steel bars should prevent it from 

widening beyond a hairline. Low stresses and natural compressive loads tend to limit 

width of cracks. Any one of a number of coating materials, including those used for 

roofing, should be sufficiently flexible and impervious to prevent leakage through such a 

small crack. 

Nonetheless, if leakage were to occur through such a crack, it might be difficult to identify the crack's location on the outside of the vessel, where it would be easiest to fix. It 

is conceivable that a leak around a penetration insert, for example, could connect to a 

crack remotely located on the inner surface of the wall. Such a situation would be similar to a roof leak, in which the location of the drip provides only a rough guide for location and repair of the failure in the roof. 

The structures team has considered several possible solutions for this issue. Low leakage 

could be assured by jacketing the concrete walls with thin (6-mm) steel plates. The 

plates would be supported by the concrete wall, and any possibility of unidentifiable or 

nonrepairable leaks would be eliminated. Simple designs are available that permit the 

welding of plates to each other and to all penetrations. Some or all of the jacket could 

be applied at any time after construction of the vessel if tests for leakage indicate that 

performance of coated concrete is inadequate. However, this adds an estimated $50,000 

to the cost and may not be necessary. 

Because no single solution has yet been selected, the team has decided to pursue preliminary design of a concrete vessel through a subcontract to an experienced structural engineering firm. As part of the design, the subcontractor will assess risks of leakage 

through the concrete walls and will recommend design and operating procedures that 

could ensure low rates of leakage. 
3.2.4 Other Major Considerations 

3.2.4.1 Position of the Condenser 

In the dome configuration, it is necessary to decide whether the evaporator should be 

located outside the condenser or vice versa. Both the radial-inflow and the axial-flow 

turbines considered for this experiment undergo lower losses of pressure when turns in 

the high-velocity steam at the turbine outlet are avoided. Thus, it is preferable to locate 

the condenser directly beneath the turbine outlet in the center of the vessel. The evaporator is best configured as an annular planform area outside of the condenser. 

3.2.4.2, Diameter of the Vacuum Vessel 

The vessel's diameter was determined primarily by the required planform area of the 

evaporator and direct-contact condenser. Limited additional area was provided for partitions, water manifolding, and access. The evaporator area was increased from the 

required minimum to provide ample clearance between the inner and outer walls of its 

partition. Additional information on heat-exchanger configuration is provided in 

Chapter 5.0 of this report. 

3.2.4.3 Height of the Vacuum Vessel 

The height of the vessel's dome was determined from the sum of four components: the 

turbinelgenerator unit, the turbine exhaust diffuser, the direct-contact condenser, and 

the discharge pool for the direct-contact condenser. The preliminary design for the 

turbinelgenerator unit established the distance between its mounting flange and the top 

of the dome at about 1.2 m (4 ft). 

The diffuser was originally specified to be 6.9 m (22.8 ft) long. Although such a long diffuser offers potentially good recovery of the energy contained in the turbine's outlet 

velocity head, it places the turbine much higher than is practical. A SERI analysis 

(Bharathan memorandum 1989) indicated that use of a diffuser 3.7 m (12 ft) long would 

reduce performance by only 2 kWe. This minimal penalty is justified by the lower cost of 

the structure and diffuser, improved access to the turbine, reduced requirements for 

crane reach, and improved frequency response of the vessel. 

A height of 1.8 m (6 ft) has been allotted for the direct-contact heat exchanger. This 

provides a 0.15-m margin over the presently specified 0.5-m ( 1.5-ft) clearance between 

the bottom of the condenser packing and the discharge pool level, 0.9-m- (3-ft)-high 

packing, and 0.3 m (1 ft) for water distribution manifolding above the packing. During 

conceptual design, designers recognized that the condenser's height could be reduced 

slightly by eliminating the margin and by positioning some of the water manifolding in 

the volume inside the inner wall of the diffuser. However, the 1.8-m height was 

maintained to provide flexibility for design modifications, should they be necessary for 

efficient operation of the condenser. 

The height of the discharge pool was fixed at 2.1 m (7 ft) based on the following 

considerations: 

Height is sufficient above the top of the cold-water discharge pipe that entrained gas 

bubbles separate from the discharge water before it enters the pipe. The pipe diameter is nominally 0.75 m (30 in.) and the pipe is located just above the foundation pad. 

Supply water for the condenser enters at grade level and turns 90" to flow upward to 

the supply manifolds above the packing. The elbow and flanges required for this turn 
must not extend above the water level of the discharge pool to ensure an unobstructed 

path for steam entering the countercurrent condenser stage. I The cold-water discharge pump is expected to have a net positive suction head (1 

between 3.7 and 6.1 m. It is preferable not to locate the inlet to this pump below the 

water table at the NPPE site (about 4 m below grade level) because excavation costs 

increase significantly below water level. 

Some flexibility in maintaining water level in this pool is desirable. 

These combined requirements fixed the overall height of the vessel's dome at 8.8 m 

(39 ft) and of the wall at 7.6 m (25 ft). 

3.2.5 Design Performance and Uncertainties 

The only factor in the structures design that affects performance of the HMTEAJNPPE is 

the rate at which air leaks into the vacuum vessel. Assuming an efficiency of 60% for 

the vent compressor, net power is decreased by about 1 kWe for each gram per secondyof 

leakage. Currently it is estimated that the combined leakage of the turbine shaft seal, 

the vacuum vessel, and all exterior piping will not exceed 1 g/s. Of this combined rate, 

i 

leakage through the vessel should be less than 0.5 g/s. This estimate was arrived at in f 

part by scaling up the HMTSTAts typical rate of leakage of 0.05 g/s. I! '2 

The uncertainty regarding this rate is 8 

based primarily on the potential for % 1 

leakage around penetrations and through 

construction joints and cracks. Flexible 

45'7YE4 sI 

exterior coatings and a steel jacket will 

ensure that the goal of 0.5 g/s can be 

achieved. 

N 

i 

Figure 3-9. Platform 

layout (a, this page) 

and penetration 

layout (b, opposite page) 

of vacuum vessel. 
3.2.6 Design Specifications 

3.2.6.1 Vacuum Vessel 

Design of the HMTEAINPPE vacuum vessel must ensure that it meets the requirements 

of other components of the apparatus. Thus, the vacuum vessel must house the heat 

exchangers, their supply manifolds and discharge pools, and the diffuser. It must support 

the turbine and some internal components and resist vacuum loads, dead weight, and 

seismic loads. It must provide access, ventilation, lighting, external viewing, and 

feedthroughs for water, pressure, power, and instrumentation wiring. Finally, it must 

limit leakage of air to less than 0.5 g/s while operating. 

The vessel achieves its size requirements by having a 6.4-m inside diameter and a 7.6-m 

height to the turbine inlet plenum flange, as shown in Figure 2-6 on page 31. 

i Requirements for piping, access, and instrumentation are met by platforms (Figure 3-9a) 

and penetrations (Figure 3-9b). The turbinelgenerator unit (including the inlet plenum 

that comprises much of the vessel's dome) is expected to weigh 36 metric tons. The 

vacuum load that the dome exerts on the vessel's walls will be 333 metric tons at 

minimum internal and maximum atmospheric pressures. 

36' penetratlon 

C.L elevation: 21.5' 

3 places 

12' penetratlon 

C.L elemtion: 21.5' 

3 places 

Platform 2. C.Lelevation: 

17.5' 

penetratlon 

elevation: 14' 

3 places 

Section A-A 

36' pipe 

C.L elevation: 25' 

36' penetratlon 

C.L elemtion: 4' 

2 places 

C.L elewtlon: 4' 

2 places 

Section B-B 

The vessel will be designed to withstand a horizontal seismic load corresponding to the 

response spectrum given in the Uniform Building Code (1988) with a peak ordinate of 

0.100 g and zero period acceleration of 0.040 g, where g is the acceleration of gravity. A 

geotechnical survey will be conducted to provide information needed to confirm or 

change the current choice of soil type 1 for the site. The vessel will also be designed to 

withstand vertical loads corresponding to the same response spectrum but with peak and 

zero period accelerations reduced to 0.067 and 0.027 g, respectively. The vertical and 

horizontal response spectra will not be applied concurrently. These specifications meet 

the general requirements for Category I (normal use) facilities. 

3.2.6.2 Control and Auxiliary Buildings 

The control building provides room and shelter for control functions, data acquisition, 

electrical switchgear, offices, and public relations. The present control building for the 

HMTSTA will provide room for extra offices, storage, and a small workshop/calibration 

area. Adjacent buildings or simple roofed structures will shelter equipment such as the 

vacuum compressor, instrument air compressor, and oil lubrication pump systems. NELH 

facilities will be available for restrooms and some inside storage. 

3.2.6.3 Crane 

In some early conceptual design work, provisions were made for a crane to be permanently installed on the site. Such a crane would be used to install and remove the steam 

turbine, the turbine inlet plenum, and equipment located inside the vacuum vessel. A 

mobile crane, column-jib crane, gantry crane, and overhead traveling crane were considered. Of these, the overhead crane was the least expensive at a cost of between $75,000 

and $125,000 (Stinnett 1989). The overhead rails could add another $50,000 to $100,000. 

An attractive alternative is to rely on mobile cranes available from local or Honolulubased construction or equipment rental companies. Although the heaviest turbine component is expected to weigh less than 15 tons, a mobile crane cannot be positioned closer 

than about 8 m from the centerline of the turbine. Thus the mobile crane would need to 

have a capacity of between 50 and 100 tons. 

Such cranes are available; one was on site for more than a year during installation of the 

I-m (40-in.) cold-water pipe system. Rental rates should not exceed $1,000 per day, 

including labor and insurance. Although the savings are significant, delays of up to one 

week might be expected before test personnel could remove the vacuum vessel's dome if 

unscheduled maintenance were required. 

During the preliminary design phase, this issue should be investigated further. Specific 

data on cost and availability should be obtained for leased, new, and surplus cranes. 

3.2.7 Layouts 

Figure 3-10 shows the general layout of the HMTEAINPPE vacuum vessel, control 

building, and auxiliary structures. Pending approval from NELH, they will be located 

between the HMTSTA and NELH1s administration building. Choice of this position is 

based, in part, on the presence of portions of the onshore seawater supply pipes. Aligning 

the vessel and control buildings with the south walls of the HMTSTA trailer and the 

NELH administration building rather than with the existing pipes is an efficient and 

aesthetically pleasing arrangement. 
The seawater supply and discharge pipes are in a corridor that permits access to 

HMTEAINPPE and HMTSTA components by forklifts and cranes. The control building 

adjacent to the vacuum vessel provides easy access for observation, wiring, and entry. A 

mobile crane can be positioned in an adjacent area to install and remove the turbine. 

The turbine, exterior wall, and foundation are the major components of the vacuum vessel (Figure 2-6). An interior wall will separate the evaporator from the condenser, 

including their respective discharge water pools. It will also support the outer wall of 

the diffuser. A grating mounted on the inner partition will support the condenser packing, the partition between the cocurrent and countercurrent stages of the condenser, and 

the inner wall of the diffuser. Columns will reinforce the grating of the condenser. 

Brackets on the inner surface of the vessel's outer wall will support the evaporator supply 

manifold and a grating in the evaporator spout area. They will also hold instruments and 

help to support the mist eliminator. 

The 1,200-sq-ft control building (Figure 3-10) houses the control room, signal wiring/ 

computer room, offices, visitors' center, electrical switchgear room, and restrooms. 

Specific construction methods and physical arrangements will be identified in the preliminary design stage. However, the building can probably be economically constructed 

using a pre-engineered (Butler-type) design. 

3.2.8 Interfaces 

The principal interfaces between the vacuum vessel and the turbine are the physical 

dimensions and loads at the turbine inlet and outlet flanges. Between the vacuum vessel 

and the seawater systems, the interfaces are the seawater pipe sizes and positions and 

the discharge pool levels. The interfaces between the vessel and the heat exchangers are 

the required planform areas, heights, and manifolding supports. In addition, steam piping 

to the surface condenser and vacuum compressor requires interfaces with the vacuum 

vessel. 

An additional "interface1' between the vacuum vessel and the turbine is necessary 

because of the desirability of completing heat- and mass-transfer tests before the turbine arrives. Because the turbine inlet plenum and rotor assembly form the vessel's 

dome, it will be necessary to ensure that parts of the turbine are delivered before the 

turbine rotor is, or that a suitable blank is available to close the top of the vacuum 

vessel. It now appears that the inlet plenum can be fabricated and delivered earlier than 

other parts of the turbine system. The rotor assembly requires the most time for 

fabrication and factory testing, and it is not likely to be available until after heat- and 

mass-transfer tests are complete. Therefore, the structures team is designing and 

planning to fabricate a suitable blank to cover the opening for the turbine rotor. 

Besides the interfaces described above between the vacuum vessel and other components 

of the HMTEAINPPE system, additional interfaces exist between other structures and 

other components. For example, platforms will be required to support the vacuum compressor, the surface condenser, and the turbine's resistive load bank. Such platforms will 

be designed to provide stable support for and adequate access to these components. Piping routes and supports, instrumentation and power cableways, and provisions for access 

of personnel and cranes require a careful interfacing between the applicable component 

design group and the structures team. 
Table 3-5. Schedule for NPPE Structures Task 

Conceptual design Jan. 1, 1989 - Oct. 31, 1989 

Preliminary design Nov. 1, 1989 - Mar. 31, 1990 

Final design April 1, 1990 - Aug. 15, 1990 

Construction July 1, 1990 - June 31, 1991 

Leak detection 

and elimination July I, 1991 - Sept. 30, 1991 

Table 3-6. NPPE Structure Costs (in $K) 

Task Equipment Contracts Consultants Total 

Preliminary design 0 

Final design 0 

Construction/testing - 680 

3.2.9 Schedule 

The schedule for design and construction of the vacuum vessel and control building at 

the time of writing is shown in Table 3-5. Unless unforeseen problems arise, the 

structures should be ready for installation of equipment by January 1991 and for heatand mass-transfer tests by July 1991. 

3.2.10 Cost and Uncertainties 

Table 3-6 summarizes estimated costs. Equipment costs include costs for site 

preparation ($loOK) and for construction of the vacuum vessel ($300K), the control 

building ($loOK), and miscellaneous structures ($30K). Equipment installation will cost 

$150K. The contract for construction and testing covers an on-site quality assurance 

inspector to test and approve forming, concrete mixture and placement, and coatings. 
CHAPTER 4.0 

SEAWATER SYSTEMS 

4.1 Objective 

The objective of the seawater systems task is to develop the seawater systems for the 

HMTEAINPPE at the STF, to manage environmental issues related to HMTEA/NPPE 

operation, and to ensure safe operation of the facility. 

4.2 Goal 

The goal of the task is to develop seawater systems for the NPPE which will consume no 

more than 81 kWe of power at the design point. Including 20 kW of power consumed in 

the NPPE heat exchangers, the goal for total pumping power is f01 kWe. The seawater 

pumps will be the largest users of parasitic power in the system. 

4.3 Approach 

The U.S. DOE has cooperated with the State of Hawaii to upgrade the capacities of the 

seawater systems serving the STF to 620 kg/s (9600 gallmin) warm seawater and 420 kg/s 

(6500 gallmin) cold seawater (DOE 1986). The seawater supply portion of this effort was 

completed and placed in limited service in July 1988. A seawater disposal trench is. 

nearing construction. The upgrade will be the basis for the HMTEAINPPE seawater 

systems. The existing portions of the upgrade are documented by Lewis, Van Ryzin, and 

Vega (1988), by Daniel (1989), and by construction blueprints from the Department of 

Accounting and General Services of the State of Hawaii (DAGS 1986). 

Adapting the STF upgrade to the needs of the HMTEAINPPE requires engineering design 

that emphasizes performance, cost, and reliability. No new research or development is 

required. During the conceptual design, the upgrade was evaluated and the potential was 

assessed for producing net power from the NPPE using the present system. It was 

determined that the existing system is inadequate to meet the goal for total pumping 

power. Design changes which will reduce the pumping power and meet the NPPE goal 

have been determined. Approaches to individual design issues are discussed below. 

Options considered and eliminated are presented, and other major considerations are 

discussed. 

Members of the seawater system team at the time of writing are Ron Davis of PICHTR, 

Lyle Genens of ANL, Herbert J. Green of SERI, and Lloyd Lewis of DOE. Ken Fong of 

PICHTR has also contributed to the conceptual design. 

4.3.1 Design Conditions 

Design conditions for the HMTEAINPPE seawater systems (Table 2-1, p. 12) were 

continuously updated and expanded during conceptual design. A complete discussion of 

these data and their sources is found in Document No. 2.1.3.6.14. 

The nominal flow rates of 620 kg/s (9600 gal/min) warm seawater and 420 kg/s (6500 

gal/min) cold seawater have been dictated by design parameters of the existing STF 

upgrade and by related environmental permits. The low and high flow rates are 50% and 

1 lo%, respectively, of the nominal flow rates. These are the projected requirements of 

the HMTEA experiments and the turbine checkout tests. The total flow rate in the 

offshore cold-water pipe, 860 kg/s (13,300 gal/min), includes the cold-water allocation 

for the HOST Park, 440 kg/s (6800 gallmin), as well as the HMTEAINPPE flow rate. 
The seawater temperatures, 26°C (78.8"F) warm and 6.1°C (43.0°F) cold, are based on 

measurements made at the HMTSTA early in 1989. The range of the warm-seawater 

temperature is the seasonal variation recorded by the NELH. The cold-seawater low 

temperature will not be known until the system can be operated at full capacity, 860 kg/s 

(13,300 gal/min), and the temperature is measured on shore. Data indicate that the 

temperature at 675 m (2,215 ft), the depth of the inlet, may actually be as low as 5.3" - 

5.8"C (41.5" - 42.4"F) (Ocean Data Systems 1977). Even if the cold seawater is below 

6.1°C (43.0°F), the remaining uncertainty is much smaller than the seasonal variation in 

the warm seawater. Thus, the nominal temperature difference, 19.9"C (35.g°F), is a reasonable design point. It is well within the range of temperature differences that will be 

available to the NPPE in the course of a year. If slightly lower cold-seawater temperatures are available, they will have the beneficial effect of increasing turbine shaft power 

without increasing parasitic power of the system. 

Data on seawater density are taken from measurements off Kahe Point, Oahu. Less than 

0.5% difference exists between the densities of the warm and cold seawater. The bar-, - 

ometric head calculations accounted for the local barometric pressure and the NPPE' ' 

operating pressures as well as the seawater densities. The variation of the force of gravity as a function of latitude was also considered. The supply and discharge barometric 

heights are essentially the same, because the change in density of seawater within the 

NPPE is quite small. The cold-water pipe density head loss of 1.0 m (3.3 ft) is from an 

experiment performed at the STF in August 1989. Previous estimates of this head loss 

ranged from 0.88 m (2.9 ft) to 1.07 m (3.5 ft) based on various projections of the seawater density vs. depth. The tides were determined from 1990 and 1991 tide tables for 

Kailua Kona harbor (NOAA 1989). 

4.3.2 Pump Configuration 

One of the most important design issues for NPPE seawater systems is configuration of 

the pump. Pump configuration refers to the location of the pumps in the system as well 

as the type of pump (e.g., submersible, vertical wet pit). The pump configuration is 

closely coupled with other design parameters. such as the head losses in the supply and 

discharge pipes and the elevations of the heat exchangers. These are discussed in Sections 4.4 and 4.5. Several configurations for supply and discharge pumps were identified 

and their relative merits evaluated. Some were less attractive and were set aside. 

A key decision reached during the first working group meeting in February 1989 was to 

use both supply and discharge pumps in the design of systems for both warm and cold 

seawater (Green Feb. 1989). This has the important advantage of allowing the heat 

exchangers to be located relatively close to grade, thus reducing overall height of the 

structure and keeping costs down. Another advantage is that net power production will 

be substantially disconnected from the tide and from variations in barometric pressure. 

For instance, the increase in pumping power of the seawater supply at low tide will be 

accompanied by a decrease in discharge pumping power. At high tide, just the opposite 

will occur. In either case, little change in net power will occur. (Observations at the 

pump station indicate that the cold-seawater sump is subject to tidal fluctuations just as 

is the warm-seawater sump. The disposal trench is also expected to respond to the 

tides.) 

Placing pumps in each supply and discharge pipe also provides the operational flexibility 

to adjust to uncertainties in head loss estimates or changes in head resulting from fouling 

of the pipes, changes in the experiment, or other causes. The result is a system capable 

of delivering the required flow rates with minimal risk. For the warm-seawater supply, a 

pump with a very low head, 1.37 m (4.5 ft), will be used so that the elevation of the 
evaporator discharge pool will be minimized (Document No. 2.1.3.6.16). A higher elevai tion would reduce the space above the evaporator needed for separation of seawater 

droplets from the steam before it enters the turbine. Because this pump will operate at 

such a low head, it will have a somewhat lower efficiency than the other pumps. However, the resulting increase in pumping power is less than l kWe. 

The impeller of the NPPE discharge pumps must be located below grade to have sufficient net positive suction head (NPSH). The configuration chosen for the discharge is the 

vertical barrel pump (Figure 4-1). The barrel provides an enclosure that operates at line 

rather than atmospheric pressure. The electric motor and the suction and discharge 

flanges all can be located above grade. Only the barrel and impeller are below grade, so 

that excavation is minimal. Location of these pumps near the HMTEAJNPPE structure 

will be convenient for both construction and maintenance. It will also minimize the 

length of the suction pipe to the pumps, thus maximizing the available NPSH. 

The configuration chosen for the supply pumps is vertical wet pit pumps located in the 

pump station sumps (Figure 4-2). This type of pump is installed in the HMTSTA seawater 

discharge sumps and is the configuration used for the OTEC-1 seawater pumps. The 

pumps have performed well in both applications. Also, this configuration has ample NPSH 

for the pumps. The cost of modifying the pump station to accommodate vertical wet pit 

pumps as shown in Figure 4-2 is reasonable, about $40K per pump (Fong Oct. 1989). 

These proposed modifications will require the approval of NELH. 

A disadvantage of this configuration is the possibility that the electric motors will be 

flooded. This could result from storm surge. In the case of the cold-seawater sump, 

flooding also could be caused by the large momentum of seawater in the offshore pipe 

following a sudden loss of pumping power or shutdown of the HOST Park pumps. For this 

reason, two backup configurations for supply pumps have been identif ied--submersible 

and vertical-barrel pumps. Further analysis will determine which configuration will be 

used. 

The submersible pump is the existing STF upgrade configuration. However, the existing 

pumps are too inefficient and deliver heads much too high for the NPPE. Further, their 

demonstrated reliability is currently inadequate for the needs of the NPPE. Progress at 

the STF with cast-iron submersible pumps and impressed current galvanic protection is 

being closely monitored. If they demonstrate improved reliability and adequate performance, they could be reconsidered as a good alternative. 

Vertical-barrel pumps can be installed in the supply pipes near the vacuum vessel. The 

potential disadvantage of this configuration is that performance of the pumps would 

degrade in response to air leakage or deaeration of noncondensable gases in the supply 

pipe coming from the pump station. This is possible because the entire pipe will be below 

atmospheric pressure while the system is operating. 

As for other configurations, vertical dry pit pumps were rejected because of the extensive modifications that would be required in the pump station. Placement of verticalbarrel pumps in the supply pipes just above the pump station was rejected because they 

could be vulnerable to high surf and would not be readily accessible for maintenance. 

Additional information about all of the pump configurations can be found in Documents 

2.1.3.6.8 and 2.1.3.6.16. 

Redundancy within the seawater systems to increase reliability also has been considered. 

For the conceptual design, single pumps have been used. The additional cost of installing 
unused pumps is not warranted for an experiment. 

During the preliminary design, three redundancy 

options will be assessed. The first will be the use of 

multiple pumps of smaller capacity, e.g., two 

I 

half-capacity pumps or three one-third-capacity 1 

pumps. If one should fail, the system at least could 

continue to operate although at a reduced level of 

performance. The second option will be an on-site 

stock of spare parts, perhaps including spare motors. 

The third will be backup pumps stored on site but not 

installed. Costs and effects on reliability of the 

I 

system will be assessed for the three options 'and 1 

recommendation will be made to the system team 1 

during the course of the preliminary design. 

4.3.3 Pump Bids 

One of the first activities for the conceptual design 

was to evaluate the possibility of using the existing 

Flygt pumps. They were found to be well oversized for 1 

the HMTEA/NPPE. They are rated to deliver .l8..9-m 

(62 ft) of head, whereas the requirements for the I 

NPPE are 3.60 rn (1 1.8 ft) and 1.37 m (4.5 ft) for the 

cold- and warm-seawater pumps, respectively (see 

Section 4.4). This evaluation also revealed that the 

impeller and motor efficiencies for submersible pumps 

are somewhat lower than those for available vertical I 

Figure 4-1. Vertical barrel 

or can pump 

Impeller 

Shaft, Pump 

Ring, Impeller 

Bushing, Bearing 

Coupling Half, Driver 

Coupling Half, Pump 

Key, Coupling 

Bell, Suction 

Seal, Mechanical, 

Stationary Element 

Nut, Shaft Adjusting 

Seal, Mechanical, 

Rotating Element 

Stuffing Box 

Collet, Impeller Lock 

Ring, Thrust, Split 

Spacer, Coupling 

Head, Surface Discharge 

Retainer, Bearing, Open 

Lineshaft 

Bowl, Intermediate 

Barrel or Can Suction 

wet pit pumps. Submersible pump impellers are typically designed with high clearances 

for wastewater service resulting in efficiencies less than 80%. The motor efficiencies 

are usually less than 90% because they are completely enclosed and must expend 

additional power to reject heat. 

An assessment of the system's performance using the existing Flygt pumps was made by 

PICHTR (June 1989). Replacing the impeller and using variable-frequency drives (VFDs) 

were investigated as means of adapting the Flygt pumps to the needs of the NPPE. The 

best combination, the existing pumps with VFD, would consume more than 101 kWe of 

power, the task goal. This assessment was based on an early configuration of the 

HMTEAINPPE. Subsequent changes have only decreased the suitability of the existing 

pumps. 

Bids were solicited from pump vendors to obtain cost and performance data for pumps 

better suited to the needs of the NPPE. Specifications for the supply and discharge 

pumps were written based on the seawater system layout as of March 1989 (Document 

No. 2.1.3.6.1 1). The specifications included warm- and cold-seawater supply pumps and 

mixed-discharge pumps. The three pump configurations specified were vertical wet pit, 

vertical barrel, and submersible. Three capacities were also included: 10096, 50%, and 

33% of the rated flow rates for both warm and cold seawater. Together, the specifications included 30 different pumps incorporating all the options under consideration at the 

time and called for cast iron or equivalent material for construction. This is the type of - 

construction used for the OTEC-1 and HMTSTA seawater pumps, all of which have performed well. 

Bids from at least 14 manufacturers were accumulated over several months. They have 

been compiled and summarized (PICHTR Sept. 1989) to provide a valuable data base of 

pump options for the NPPE. Although the seawater system layout has changed since the 

specifications were written in March, the bids indicate the prices and performance 

expected from the pumps that eventually will be purchased. Many were quoted with high 

impeller efficiencies of 80% to 89%. Several brands of barrel pumps are available that 

have the required NPSH between 1.8 and 6.1 m (6 and 20 ft). This is an important characteristic for the discharge pumps, as noted previously. Prices for the 100%-capacity 

pumps were generally between $20~ and $40K per pump for the sizes appropriate to the 

current configuration of the system. Curves shown in Figures 4-3 and 4-4 are examples 

of the performance of the pumps. Operating points for all four pumps are noted on the 

curves. Efficiencies range from 76.5% for the warm-seawater supply pump to 85.5% for 

the cold-seawater supply pumps. 

Most of the bids were for pumps with mixed-flow or axial-flow impellers. They have two 

inherent characteristics that set them apart from most centrifugal pumps. The shutoff 

head is usually significantly higher than the head at the design point, which creates the 

possibility of overloading the motor. They would require safety features such as alarms 

for high motor temperature, high current draw, or high discharge pressure. In addition, 

they typically have areas of unstable operation at low rates of flow. The preliminary 

design will include careful evaluation whether operating conditions might occur in which 

potential exists for these instabilities. 

4.3.4 Electric Motors 

In vertical-shaft pumps, the thrust bearings are generally put in the electric motor. The 

pumps themselves have only shaft bearings, so that the weight and thrust loads of the 
Warm Seawater Supply Pump: Warm Seawater Discharge Pump: 

9600 gprn (620 kgls) 9600 gprn 1620 kgls) 

4.5 ft head (1.37 m) 10.5 ft head (3.20 m) 

76.50h efficiency 81.5% efficiency 

16 ft NPSH (4.88 m) 16 R NPSH (4.88 m) 

U.S. Gallons per Minute 

I I I I I I I 

200 300 400 500 600 700 800 

kgls of Seawater 

Figure 4-3. 

Sample warm-seawater 

pump curves 

shaft and impeller are suspended from the motor thrust bearing. The height of the 

impeller in the bowl is determined by an adjustment in the coupling between pump and 

motor shafts. Typically, the pump and motor are shipped to the site separately because 

of the weight of the motor. The unit is assembled on site and the impeller height is then 

adjusted. 

High-thrust vertical motors have two configurations: solid shaft and hollow shaft. In vertical hollow-shaft motors, the pump shaft extends through the motor and is coupled to 

the motor shaft on top. In vertical solid-shaft motors, the shaft coupling is below the 

motor. The two domestic suppliers of high-thrust vertical electric motors are General 

Electric and U.S. Motors. 

Performance of electric motors is affected by the type of enclosure. The more completely enclosed and weatherproof a motor is, the less effectively it rejects heat. 

Totally enclosed fan-cooled (TEFC) motors require a fan for cooling, which reduces efficiency. Open drip-proof motors may be 1% or 2% more efficient than comparable TEFC 

motors. For outdoor installation, TEFC motors are recommended. General Electric 

offers premium-efficiency TEFC motors with efficiencies between 90% and 94% in the 

sizes of interest for the HMTEA/NPPE, 15 to 37 kW, (20 to 50 hp). Certification of 

performance by factory testing is available for an additional fee. 

The motor specifications should include materials and coatings appropriate for a nearocean location (severe duty). Auxiliary heaters inside the motor to keep it dry should 

also be specified. Such heaters have effectively prevented condensation of atmospheric 

moisture inside motors used in humid environments. For applications with VFDs, highertemperature insulation such as Class F is recommended, because a motor rejects heat 

less effectively as it is slowed down. Class B is adequate for single-speed motors. 

Purchasing the pump, motor, and VFD as a package may be advantageous because it 

centralizes responsibility for performance and service. 

Motors are rated for the power required by the pump and must not overload over the full 

operating range of the pump. Motors are rated for an altitude of 1000 m (3300 ft). At 

sea level, a given motor will cool itself more efficiently which, in effect, provides some 

factor of safety above its nameplate power rating. Motors typically come with 1.0 or 

1.15 service factors. A service factor of 1.15 is equivalent to excess rating of 15%. For 

a power system with nominal voltage of 480 volts, the motor nameplate voltage will be 

460. 

4.3.5 Control of Flow Rate 

Variable-frequency drive (VFD) has been selected to control the flow rates in the NPPE 

seawater systems. VFD is typically more energy efficient than a throttling valve or a 

bypass loop for controlling flow over a range of flow rates and heads. The change in head 

or flow need only be greater than the drop in efficiency caused by addition of the VFD to 

the electrical circuit. Typical VFD electrical efficiencies are about 95%, so any change 

in either flow or head or both that is greater than 5% will make VFD the most energy 

efficient method. When operating at synchronous speed (e.g., 60 hz), the rectifiers in the 

VFD can be bypassed so that it is virtually 100% efficient. 

VFD also offers greater operating flexibility than throttling valves. The seawater systems design requires a range of flow rates from 50% to 110% of the nominal flow rate 

and a range of tides of -0.37 m (-1.2 ft) and +0.61 m (+2.0 ft). The discharge pumps may 

operate against a head that increases with time if degradation should occur in disposal 

trench performance. Furthermore, the estimates of the head losses in the piping runs 

have some uncertainty. A pump controlled with a VFD can adjust to these kinds of 

changes. The VFD can drive the pump above the synchronous speed of the electric 

motor, allowing the pump to deliver heads, flow rates, or combinations of the two that 

are greater than those originally specified. It is only necessary that the VFD and the 

motor be sized for the increased electrical loads. 

Another advantage of VFD control of flow rate is that pumps may be soft started, allowing operators to start the seawater flow in a gradual, controlled manner. This will be 

especially helpful because levels of seawater at the pump station and disposal trench are 

functions of the flow rates through them. As those levels change during start-up, shutdown, or other operational transients, the required pump heads will also be changing. 

Finally, soft starts prevent large inrush of current to the electric motors. The selection 

of VFD to control flow rate is discussed in Document No. 2.1.3.6.12. 

4.3.6 Supply and Discharge Pipes 

Onshore supply pipes for both warm and cold seawater were installed between the pump 

station and NELH during construction of the STF up rade. Both pipes are 0.61-m (24-in.) 

high-density polyethylene (HDPE) pipe about 180 m 7 600 ft) long. Sections of these pipes 

are fully buried to prevent damage from high surf and to pass beneath roadways. For 

most of their length, however, the pipes are buried only to their centerline. Their layout 

is described in Section 4.4. 

Because HDPE is flexible, the pipes are subject to collapse under internal vacuum. 

Analysis of the STF supply pipes, performed with the assistance of the manufacturer, 

determined that the buried sections would withstand the vacuum that would exist during 
operation but the half-buried sections would not. The buried pipe performs better 

because the backfill around it helps to maintains its circular cross section. This analysis 

is presented in detail in Document No. 2.1.3.6.9. For the HMTEAINPPE, the supply pipes 

must either be replaced or modified with some type of mechanical restraint. Suitable 

restraints might include a gravel berm over the pipe, a poured concrete cover, or steel 

reinforcing bands. 

During the first working group meeting in February 1989, it was decided that the existing 

warm-seawater supply pipe would be replaced with pipe of larger diameter to reduce the 

head loss (SERI March 1989). A 0.91-m (36-in.) pipe is recommended which will reduce 

the head loss about 1.68 m (5.5 ft) at 640 kg/s (9600 gallmin) which in turn will reduce 

pumping power nearly 15 kWe. The basis for this estimate is found in Section 4.5. 

Alternative materials that have been suggested for the new pipe are buried HDPE, fiberreinforced plastic, PVC, coated steel, concrete-lined steel, and uncoated steel. These 

materials are being evaluated. The important criteria are capability to withstand 

vacuum, leaktight joints, resistance to corrosion, and installed cost. Selecting the 

material is a top priority for the seawater systems team during the preliminary design. 

A route for installation of the pipe must also be determined. The option of removing the 

existing 0.61-m (24-in.) pipe and replacing it with the 0.91-m (36-in.) pipe is under 

consideration and must be approved by NELH. 

A decision is also pending on the cold-seawater supply pipe. The existing 0.61-m (24-in.) 

pipe will have head losses of only 0.98 m (3.2 ft) at 420 kgls (6500 gallmin), requiring 

pumping power of 5.4 kWe. Thus, the incentive is only moderate to replace the pipe 

simply to increase the diameter. If the pipe is replaced, a 0.76-m (30-in.) pipe would be 

adequate, reducing the pumping power to 1.4 kWe. 

The existing pipe could be modified to withstand vacuum as previously noted. There are 

concerns, however, about the integrity of numerous flanged joints in the pipe run, many 

of them buried and therefore inaccessible. Each is a potential source of leaks. A leak 

check has not yet been done because an NELH tenant, Ocean Farms Hawaii, is continuously dependent on this pipe for cold water. Additional sources of supply may make this 

pipe available for DOE use during 1990. Further analysis and testing will be done to 

enable a decision about this pipe during the preliminary design. 

Pipes will be needed to connect the HMTEAINPPE to the discharge pumps and the 1.2-m 

(48-in.) discharge pipe. The material chosen for the warm-seawater pipe will also be 

used for these. A 1.2-m (36-in.) warm and 0.76-1-11 (30-in.) cold discharge pipe from the 

HMTEAINPPE vacuum vessel to the discharge pumps will be located about 12.2 m (40 ft) 

away. A 0.91-m (36-in.) mixed-discharge pipe will run about 36.6 m (120 ft) from the 

pumps to connect to the 1.2-m (48-in.) pipe that runs to the trench. 

4.3.7 Concrete Sump 

A concrete disposal sump was included in the STF upgrade completed in 1988. Its purpose 

was to mix seawater discharged from the STF with seawater bypassed from the supply 

pipes going to the STF. This sump has been carefully evaluated and found to be inadequate for use in the HMTEAINPPE. 

For the HMTEAINPPE, the sump would be required to operate at a pressure of one-half 

atmosphere. Structural analysis showed that the sump's flat cover would not bear the 

loads imposed by this internal vacuum. In addition, the 1.2-m (48-in.) pipe penetrations 

to the sump probably are not currently vacuum tight and cannot be easily modified to be 
so. Using the sump in the discharge system would add a source of leaks as well as a high 

point which would require venting with a vacuum pump. Additional head losses would 

also occur in the seawater flow into and out of the sump. 

Finally, the possibility was investigated of using the sump as a bypass between the 

HMTEAINPPE supply and discharge pipes. It was found that pressure in the sump will be 

higher than that in the seawater supply pipes. If the bypass valves to the sump were 

opened, seawater would flow the wrong direction, from the sump into the low-pressure 

supply lines. Given all of these findings, the concrete sump was eliminated from the 

NPPE design. 

To provide the capability of bypassing the sump, an additional pair of flanges will be 

added to the 1.2-m (48-in.) discharge pipe. They will be located 7.3 m (24 ft), 6 pipe 

diameters, from the first flange at the concrete sump. This will create a spool piece 

that will later be removed (Figure 4-5) for making the connection to the HMTEAINPPE. 

In the interim, the concrete sump will be available for flowing seawater to the disposal 

trench for tests. 

4.3.8 Disposal Trench 

The State of Hawaii, through its contractor R. M. Towill Corporation, is responsible for 

design and construction of the disposal trench and 1.2-m (48-in.) HDPE discharge pipe. In 

July 1989, Towill submitted a final design in the form of a Prefinal Bid Document for 

DOE review. It covers the first phase of construction, 30.5 m (100 ft) of trench, and 

installation of the 1 -2-m (48-in.) pipe. Towill incorporated review comments into a 

Revised Prefinal Bid Document submitted in September. This document, with minor 

modifications, was promptly approved by DOE as the final design of the trench. The 

seawater systems team supported DOE throughout this process by writing design specifications and providing technical review of the bid documents. The state is proceeding to 

award a construction contract. The trench should be complete and ready for testing by 

mid-1990. 

Com lete details of the final design are available from the bid document (R. M. Towill 

1989f The locations of the pipe and trench can be seen in the site layout, Figure 2-9 

(p. 39). Numerous features have been incorporated to meet HMTEAINPPE 

requirements. The 1.2-m (48-in.) pipe will slope continuously down to the trench. No 

sharp turns or fittings will create additional pressure drop or trap air bubbles. A spool 

piece added to the pipe as noted above (Figure 4-5) will accommodate bypass of the 

concrete sump. The entire pipe will be buried 1.2 m (4 ft) deep. Thus constrained, it will 

be able to withstand internal vacuum without collapse. The heavy-walled HDPE pipe 

available at NELH, DR21, will be used, which will also provide greater resistance to 

collapse. The top of pipe will be at least 0.3 m (I ft) below the lowest tide as it enters 

the trench. The elevation of the bottom of the trench will be surveyed to expedite 

measurements of the water level. 

Geolabs-Hawaii, designer of the trench, estimated a capacity between 65 and 260 kg/s 

(1000 and 4000 al/min) for the first 30.5 m (100 ft) of trench based on permeability tests 

done in 15-cm f 6-in.) wells drilled at the trench site. Their estimate allowed for 3.0 m 

(10 ft) of head in the trench. Given that the dischar e flow rate from the HMTEAINPPE 

is 1040 kg/s (16,100 gallmin), more than 122 m f 400 ft) of trench may by needed. 

Furthermore, a 3.0-m (lo-ft) head in the trench would consume more than 40 kW of 

pumping power, which would seriously affect net power production. (See Section $.?for 

details of performance projections.) Because of the characteristic inhomogeneities of 

the lava rock at Keahole Point, the actual performance of the disposal trench could be 
Finish Grade Fill to Provide 

Along Pipe 4" Min. Pipe Cover 

Ex. Ground 

Ex. Sump Box To Disposal Trench 

Ex. 48" HDPE Pipe 

Existing and New 

48"@ HDPE Pipes 

Figure 4-5. Concrete sump and 1.22-rn (48-in.) discharge pipe 

with spool piece (R.M. Towill 1989) 

significantly different than Geolabs has estimated. When the first 30.5 m (100 it) of 

trench is in and tested, better estimates of its performance and of net power production 

will be made. 

During the conceptual design, no other disposal methods have been investigated. During 

the preliminary design, alternatives to trench disposal should be assessed, taking into 

account both cost and environmental regulations. 

The NPPE baseline design from 1987 specified a 1.8-m (6-ft) discharge head dictated by a 

weir in the trench. During the first working group meeting in February 1989, the weir 

was eliminated. Subsequently, the seawater systems team has used an optimistic goal of 

0.9 m (3 ft) of head in the trench as a basis for obtaining bids from pump vendors and 

estimating performance of seawater systems. Net power production will decrease about 

4.3 kWe for each additional foot of head. 

4.3.9 Seawater System Tests 

Flow tests are planned for the STF upgrade of the seawater systems including the disposal trench. The data from these tests will be considered in the design of the seawater 

systems for the HMTEAINPPE. The data are especially important because the seawater 

systems will be the major consumer of parasitic power. 

The goal of the flow tests is to measure the hydraulic performance of the seawater systems over a range of operating conditions. Characteristics to be measured include head 

losses in both the onshore and offshore pipes, sump operating levels at the pump station, 

seawater temperatures, and elevation of the disposal trench pool. These data are particularly critical for selecting pumps for the NPPE that have appropriate head-vs.-flow 

characteristics, sufficiently low NPSH requirements, and high operating efficiency. 

These tests have not yet been performed because an adequate disposal system for the 

seawater is lacking and because of use of the cold-seawater pumps and supply pipe by 

Ocean Farms Hawaii, as noted previously. Limited tests were performed in August 1989, 

when NELH provided an hour during which the cold-seawater flow rate could be varied 

and head losses measured. Similar additional tests and flowmeter calibrations are 

pending, 
A draft test plan for a full set of seawater system tests has been written during the conceptual design. This plan will be promptly completed during the preliminary design, so 

that instruments can be obtained and calibrated and testing can begin as soon as the disposal trench and cold-seawater supply pipe become available. The discharge trench and 

associated 1.2-m (48-in.) discharge pipe are currently scheduled for construction in early 

1990 (see Section 4.3.8). The tests are also contingent on environmental monitoring that 

is already under way (see Section 4.3.10). For these tests, seawater will be pumped from 

the pump station through the existing 0.61-m (24-in.) pipes to the concrete disposal sump 

and on to the disposal trench. Construction of the trench and completion of the subsequent flow tests is critical for the final design. The preliminary design can proceed 

based on head-loss calculations alone. 

4.3.10 Environmental Issues 

Environmental impact statements were prepared by the High Technology Development 

Corporation (1985) and the Natural Energy Laboratory of Hawaii (NELH 1987) for the 

STF seawater system upgrade. DOE (1988) has completed a "finding of no significant 

impact" for the system. These documents allow construction and operation of the system 

to proceed. The potential impacts of disposing seawater into the trench were found to be 

either reversible or readily mitigated. Environmental monitoring is mandated to enable 

early detection of any unacceptable impacts. If such impacts are detected, appropriate 

corrective action will be taken. Reaeration of the discharged seawater will be investigated as a means of mitigating potential impacts of discharging oxygen-depleted 

seawater. 

The monitoring will be part of a broader program developed to meet the needs of all 

users of seawater at NELH and the HOST Park. A draft plan for this program, called the 

Cooperative Environmental Monitoring Program (CEMP), was circulated to interested 

parties for review in April 1989. The seawater systems team, through DOE, provided 

comment to NELH. The team found that the CEMP was adequate to provide monitoring 

before and during operation of the disposal trench and the HMTEAINPPE. DOE is one 

source of funds for this program. 

Although the CEMP has not been completed and published, NELH has begun sampling the 

groundwater, the near-shore seawater, and the anchialine ponds. These samples will provide the baseline data against which future data will be compared when the trench is 

operating. By the time the trench is completed in 1990, sufficient baseline data should 

be available so that testing and operation will not be delayed. 

4.3.1 1 Plant Safety 

In keeping with the task plan, no formal activities with regard to plant safety were conducted during the conceptual design. However, plant safety and the possibility of failures have been considered by all of the design teams. Structured activities regarding 

plant safety will begin during the preliminary design phase, and analysis of failure modes 

and their effects for the entire HMTEAINPPE system will be completed. Safe operating 

procedures will be drafted to provide specific guidance for the final design. 

4.4 Lavout 

A process flow sheet for the NPPE including the seawater systems is found in Figure 2-7 

(p. 32). The site plan in Figure 2-9 (p. 39) shows the overall layout for the onshore 

sections of the seawater systems, including routes of the existing 0.61-m (24411.) supply 

pipes and the future locations of the 1.2-m (48-in.) discharge pipe and disposal trench. 
The route for the replacement 0.91-m (36-in.) warm-seawater supply pipe will be 

determined during the preliminary design. The supply pipes are shown passing through 

the existing flowmeter enclosure and on to the HMTEAINPPE. Use of these flowmeters 

also will be evaluated during the preliminary design. 

Design of the seawater systems has been coordinated with that of the HMTEAINPPE 

structure to determine the elevations of the evaporator and direct-contact condenser. 

These elevations, with the relevant heads and barometric heights for the seawater 

systems, are shown schematically in Figures 4-6 and 4-7. Discharge pumps have been 

incorporated into the design so that the heat exchangers can be located relatively close 

to grade rather than at the barometric height. This reduces the overall height of the 

HMTEAINPPE structure, which will help keep costs down. 

The design includes both supply and discharge pumps (see Section 4.3.2). This configuration provides confidence in the system's function and the operating flexibility that is 

necessary for an experiment. Inclusion of a warm-seawater supply pump places pool elevation of the evaporator 4.1 m (13.5 ft) above grade. At this elevation, ample headroom 

is available above the evaporator for droplets to separate from the steam before the 

steam enters the turbine. 

The overall height of the HMTEAINPPE structure is determined by the elevation of the 

direct-contact condenser's discharge pool, the height of the direct-contact condenser, 

and the length of the turbine diffuser. Thus, if the pool elevation can be reduced, the 

overall height will be reduced. However, the pool must be maintained at an elevation 

that provides adequate NPSH for the discharge pumps. The vertical-barrel pumps that 

will be used to discharge seawater will have impellers below grade to gain NPSH. 

However, construction experience with the HMTSTA sumps and with the STF upgrade 

pump station indicates that the excavation of pits for these pumps below the 

7 HEAD LOSSES 

IN ON-SHORE 

\ T f 1 . . 

CONFIGURATION: 36' n PIPE C. 

VERTICAL WET PIT 

WW SUPPLY PUMP ELEVATION OF EVAPORATOR 10.5' 

0.9irn 36in.) O.D. f25.6 

STEEL SUPPLY PlPE 

HEM LOSS IN 

DESIGN CONDITIONS, EVAPORATOR SUCTON PIPE 

3.3' 0.7' 

BAROMETRIC 

620kg/s (9600gpm) 13 s- HEIGHT 

SEPARATELY PUMPED I 

DISCHARGE BAROMETRIC 

HEIGHT 

324' 

I 

1 3' 

HEAD LOSS 

OFF-SHORE 

PlPE -0.1' 

4.6' 

YMI SUPPLY Pumping Power: DISCHARGE TRENCH 3.0' 

WMP HEAD 

4.5' Supply - 128 off - 88 

Dbcharge - 27.1Ieff:- 72353 

39.7kw 

Figure 4-6. Warm-seawater system heads and elevations, draft, 9/22/89 (rn = ft x 0.3048) 
HEAD LOSS IN 

DISCHARGE PIPE-) 

0.9' 

HEM LOSSES 

M ON-SHORE 

24' PlPE 1 

CW DISCHARGE 

PUMP HEM 

17.7' 

DENSlTY HEM LOSS 

3.3' Figure 4-7. Cold-seawater system heads and elevations, draft, 

11/08/89 (m = ft x 0.3048). Note that static head loss in 

DCC is higher than nominal value. 

I HEAD I ELEVATION OF DCC 

BAROMETRIC IN DCC 6-8' 

CONFIGURATION: HEICHT DISCHARGE POOL 

327' 

MRllCAL WET PIT +18.9' HEIGHT 

CW SUPPLY PUMP SUCVON PIPE 327' 

EXISTING 0.61~1 (24in.) 

SUPPLY PlPE 

DESIGN CONDlllONS 420kg/s 

(6500gpm) 

groundwater level is time consuming and expensive. Further, by staying above the 

groundwater level, the need for a watertight liner for the excavation is avoided and the 

potential is reduced for corrosion on the pumps' exterior surfaces. 

SEPARATELY PUMPED 

DISCHARGE 

The groundwater levels at the HMTEAINPPE site are estimated to be between 2.7 and 

3.6 m (9 and 11 ft) below grade at high tide. Additional data will soon be available from 

groundwater sampling wells monitored by NELH. Allowing for about 0.3 m (1 ft) head 

loss in the suction pipe, the direct-contact condenser's discharge pool will be as much as 

5.2 m (17 ft) above the groundwater at high tide. This clearance should be adequate. 

Several vertical-barrel pumps have been identified with required NPSH of 4.6 m (15 ft) or 

less. 

13' 

I M SI 

4.5 Performance 

-0.4' 

HEAD LOSS DISCHARGE TRENCH 

Purnplng Power. 3.0' 

51.5kw 

To predict the performance of the seawater systems, conservative projections for efficiency of the components have been made on the basis of information from vendors. An 

efficiency of 95% was estimated for the VFDs and 92% for the electric motors. An efficiency of 82% applies to both discharge and cold-seawater supply pumps. A lower efficiency, 7596, has been projected for the warm-seawater supply pump because it must 

deliver a very low head and so will operate at less than full efficiency. 

Combining projections for pump, motor, and VFDs, the overall efficiency for the three 

high-efficiency pumps is 72%. The warm-seawater supply pump will operate at 66% 
overall. Because this pump is delivering a very low head, 1.4 m (4.5 ft), the penalty for 

the lower efficiency is less than 1 kW . These efficiencies and other performance 

parameters are summarized in Table 4-f. There are at least two opportunities to get 

better performance than is projected in the table. First, vendors indicate that 

efficiencies of 85% or better may be obtained from the seawater pumps. Second, if the 

pumps can be operated at 60 Hz, then the VFD can be bypassed so that no decrease in 

ef fici'ency occurs. 

Head losses in the seawater pipes have been estimated using standard methods. Friction 

factors have been taken from the Moody diagram using pipe roughness of 0.015 mm 

(0.00005 ft) for clean pipe as recommended by the manufacturer. K-factors for minor 

losses in fittings and valves have been taken from Crane (1986). The head loss from the 

pool elevation in the disposal trench above sea level has been estimated as 0.9 m (3 ft). 

The uncertainties in this pool elevation are discussed in Section 4.3.8. The head losses in 

the heat exchangers have been estimated by the heat exchanger team. The head losses 

calculated for the individual pipe runs are noted in Figures 4-6 and 4-7 and Table 4-1. 

Details of these calculations are found in Document No. 2.1.3.6.19. The head losses will 

be measured to add confidence to the design as described previously in Section 4.3.9. 

The pumping power is estimated to be 39.7 kW for warm seawater and 51.5 kW for cold 

seawater, for a total of 91.2 kWe--almost 15 kW below the goal of 101 kte. The 

sensitivities to changes in head have been estimate%by differentiating the expression for 

pumping power. Additional head will result in an additional 8.5 kWe/m (2.6 kWe/ft) of 

pumping power for warm seawater, 5.6 kW /m (1.7 kWe/ft) for the cold seawater, and 

14.1 kW,/m (4.3 kWe/ft) for the combined fkws. The uncertainty in the pumping power 

calculation is estimated to be &lo%. A 1% change in overall pumping efficiency will 

change the warm-seawater pumping power by 0.54 kW , the cold-seawater pumping 

power by 0.71 kWe, and the combined flows by about 1.25 kWe. The seawater system's 

insensitivity to tides or to changes in barometric pressure was d~scussed in Section 4.3.2. 

4.6 Specifications 

Specifications for the HMTEAJNPPE seawater systems are summarized in Table 4-2. 

Supporting information is found in Tables 2-1 and 4-1 and Figures 4-6 and 4-7. 

4.7 Interfaces 

3 

The interfaces between the seawater systems and the rest of the HMTEAINPPE are 

shown in Figure 4-8. Various electrical services are required, and compressed air will be 

used to operate valves. Control signals will be received from the control room, and 

instrumentation signals will be sent to the control room. The physical interface with the 

HMTEAJNPPE structure itself will be at flanges outside the vacuum vessel. All plumbing 

inside the vessel will be designed in the heat exchanger task. The pumps and piping for 

the miscellaneous cooling loads will be designed in the ancillary systems task. Pumps and 

piping for the desalinated water condenser will be designed by the heat exchanger team. 

The seawater supply interface is the ocean, and the seawater discharge interface is the 

disposal trench. 

4.8 Schedule 

A proposed compressed schedule for the seawater systems task at the time of writing is 

shown in Table 4-3. This schedule provides for the seawater systems to be operational on 

July 1, 1991. The first phase of trench construction must be finished before flow tests 

can proceed. Data from the tests are needed for the final design and to determine what 
Table 4-1. NPPE Seawater Systems Performance Summary (as of 11-15-89) 

Warm-Seawater System (at 620 kg/s, 9600 gallmin) 

Offshore pipe head loss 

Onshore pipe head loss 

Evaporator static loss 

Evaporator dynamic loss 

Discharge pipe head loss 

Discharge trench head loss 

Total warm-water head loss 

Supply pump head 

Overall pumping efficiency 

Supply pump power 

Discharge pump head 

Overall pumping efficiency 

Discharge pump power 

Total warm-water pumping power 

Cold-Seawater System (at 420 kg/s, 6500 gallmin) 

Offshore pipe head loss 

Onshore pipe head loss 

Cold-water density head loss 

Condenser static loss 

Condenser dynamic loss 

Discharge pipe head loss 

Discharge trench head loss 

Total cold-water head loss 

Cold-water supply pump head 

Overall pumping efficiency 

Cold-water supply pump power 

Cold-water discharge pump head 

Overall pumping efficiency 

Cold-water discharge pump power 

Total cold-water pumping power 

Component Efficiencies 

Variable-f requency drives 

Electric motors 

Warm-water supply pump 

All other pumps 
Table 4-2. Seawater Systems Specifications as of 11-15-89 

Warm Seawater Cold Seawater 

Flow rates 

Supply pipe size 

HMTEAINPPE flange sizes 

Supply 

Discharge 

Discharge pipe size 

Evaporator pool elevation 

Direct-contact condenser 

pool elevation 

Disposal trench head 

Mixed-discharge pipe sizes 

36 in. 24 in. 

0.91 m 0.61 m 

30 in. 24 in. 

0.76 m 0.61 m 

36 in. 30 in. 

0.91 m 0.76 m 

36 in. 30 in. 

0.91 m 0.76 m 

13.5 ft MSL 

4.1 m MSL 

7.0 ft MSL 

2.1 m MSL 

36 in. (0.91 m) from discharge 

pumps to the STF upgrade 

discharge pipe; 48 in. (1.2 m) 

to the disposal trench 

additional length of disposal trench is needed. A second phase of trench construction 

shown on the schedule is less critical. 

4.9 Cost Estimate 

A cost estimate for completion of the HMTEAINPPE seawater systems is given in 

Table 4-4. The estimate for onshore piping includes replacing existing warm-seawater 

supply pipe with a 0.91-m (36-in.) pipe, extension of the existing cold-seawater supply 

pipe to the HMTEAINPPE, and discharge pipes to connect the HMTEA/NPPE to the 

1.2-m (48-in.) discharge pipe. The cost to replace the existing 0.6 1-m (24-in.) coldseawater supply pipe is noted separately because it is only a design option at this time. 

The cost of the piping is based on quotes for coated steel pipe obtained during the 

conceptual design. The cost of installation is scaled from the construction cost estimate 

done by R. M. Towill for the 1.2-m (48-in.) HDPE pipe. 
L 

0 

6 oT < -- - > >gs c 0 

OoO 0 0 .w 

792 m 0 o Return 

Interfaces Interfaces 

rn rnc urn 1 Evaporator r----1 I - - 

.g 8 5 Valve Positions 7 - I 

I 0 Sump Levels I I I Direct -Contact 

Condenser I Flow Rates _, I 

Motor Speeds Desalinated Water 

I 

I Current I 

Condenser 

I I 

I - Voltage 

_J Vacuum Compressor Precooler 1 

I 

I Power Factor - Pressure 

I 

I I - 

Temperature ther Cooling Needs I 

I -------- J 

I 

I 

1 Reaeration -c I 

I I 

I - I 

I 

I 

I 

1 

I 

Figure 4-8. Seawater systems interface diagram, 12/6/89 

Table 4-3. Seawater Systems Schedule (as of 11-15-89) 

Preliminary design October 1, 1989 - February 28, 1990 

Trench construction (Phase I) January 1, 1989 - February 28, 1990 

STF flow tests March 1, 1989 - March 31, 1989 

Final design March 1, 1989 - June 30, 1990 

Trench construction (Phase 11) August 1, 1990 - October 31, 1990 

and testing 

Purchase pumps July 1, 1990 - March 31, 1991 

Install onshore pipes October 1, 1990 - June 30, 1991 

Install and check out pumps April 1, 1991 - June 30, 1991 

Seawater systems operational July 1, 1991 
Table 4-4. Cost Estimate for the HMTEAINPPE Seawater 

Systems ($K) (as of 12-5-89) 

Subtotals Totals 

STF flow tests 

a instrumentation 

A&E consulting firm 

drawing preparation 

Safety consultant 

Warm-seawater supply pipe 

(replace existing 24-in. pipe) 

purchase 

a install 

Cold-seawater supply pipe 

(extend existing pipe) 

a purchase 

install 

Discharge pipes 

a purchase 

a install 

Seawater pumps 

warm-water supply 

a cold-water supply 

a warm-water discharge 

a cold-water discharge 

installation 

($40K each) 

a VFDs 

Instrumentation 

Task total 

POTENTIAL ADDITIONAL EXPENSES: 

Disposal trench 

a additional 330 ft. 

Replace 24-in. cold-water supply pipe 

purchase 

install 

TOTAL: 
Costs of seawater pumps are based on the bids recently received from vendors for castiron pumps. Installation costs for vertical wet pit supply umps in the pump station are 

based on an itemized cost estimate (PICHTR Sept. 1989 f that was in good agreement 

with a similar estimate by R. M. Towill. Installation costs for the discharge pumps were 

estimated to be the same as those of the supply pumps. More detailed assessment of this 

cost is needed during the preliminary design. 

The cost for the first phase of trench construction is not included in Table 4-4; that 

money was previously set aside by DOE. The cost of extending the trench is $1 lOK per 

30 m (100 ft), as noted at the bottom of the table. The cost of the extension is projected 

from R. M. Towill's estimate for the cost of the first 30 m (100 ft). 
CHAPTER 5.0 

HEAT EXCHANGERS 

5.1 Objective and Goals 

The objective of the heat exchanger task is to develop the heat exchanger hardware 

required for the HMTEAINPPE, including subsystem interfaces to other system 

components, with minimal technical risk to assure net power production from the overall 

system. 

The components originally assigned to the heat exchangers task are 

Evaporator 

Direct-contact condenser 

Surface condenser 

Containment vessels 

Predeaerators 

Mist eliminator 

Water distributor 

Steam distributor. 

The components considered under this task, with the interfaces to other subsystems, are 

shown in Figure 5-1. 

The heat exchanger components developed for the NPPE will be tested during the 

HMTEA phase of the project to confirm performance predictions and to minimize 

technical risk for the NPPE tests. Results of the HMTEA tests will also be used to 

validate the predictive codes developed for the heat exchanger components at a much 

larger scale than tested in the HMTSTA. 

The goals of the task, through the conceptual design phase, were as follows: 

I. Complete all tests on the direct-contact condenser at the HMTSTA using seawater, 

at operating conditions near those originally planned for the NPPE. 

I 2. Analyze the results obtained for direct-contact and surface condensers. 

3. Complete comparisons to existing models for the direct-contact and surface condensers and document their verification so that they may be used in the design 

effort. 

4. Complete the conceptual design of heat exchanger components for the 

HMTEA/NPPE and provide the necessary documentation to proceed with the next 

step in the design effort. 

All goals have been attained; this document summarizes the background of the design, 

the issues considered and their resolution, and the details of the design that the heat 

exchanger team developed. 

5.1.1 Composition of the Heat Exchanger Task Team and Documentation 

The heat exchanger team, formed at the inception of the HMTEA/NPPE design effort in 

February 1989, is composed of Federica Zangrando, SERI, task leader; Steve Masutani, 

PICHTR; and C. 8. Panchal, ANL. Additional members added to the team during the 

conceptual design period are Ken Fong of PICHTR and Tom Rabas of ANL. 

The task members discussed issues related to the task in meetings held on March 2, 1989, 

and April 13, 1989. An additional meeting was held with PICHTR personnel on 

October 24, 1989. Task members also communicated by telephone and in writing. 

The discussions are documented in SERI memoranda (SERI reference document numbers 

2.2.3.2, 2.2.3.3, 2.2.3.4, 2.2.3.5, 2.2.3.8, and 2.2.3.9) written between February 28, 1989 

and October 26, 1989. Two letters from ANL and two from PICHTR supply additional 

details regarding options, analyses, and design calculations. These items are available 

from any task member. Details provided in these internal documents are referred to in 

the text. The cited documents are listed in the references (Chapter 8.0). 

5.2 Approach to Design 

The major considerations in the design of the HMTEAINPPE heat exchangers are 

r to keep the design structurally and functionally simple 

r to provide for minimum maintenance and operational requirements 

r to employ demonstrated performance characteristics of the major components. 

The design must meet the operational requirements of the NPPE to ensure that net 

power will be produced. If additional flexibility can be built in that interferes neither 

with NPPE operation nor with the considerations above, then the design should provide 

for possible modifications so the HMTEAINPPE apparatus will be more suitable for 

additional developmental tests after net power goals have been attained. 

5.2.1 Baseline and Nominal Operating Conditions 

Design conditions for the heat exchangers have been updated since the Dillingham/Lucas 

design of the HMTEA. The development of the NPPE design and of the operational 

parameters is discussed in Section 2.3 and summarized in Table 2-3. The parameters 

pertinent to design of the heat exchangers are summarized here. 

Baseline conditions for warm- and cold-seawater flows available for the NPPE have 

remained 9600 gpm (620 kg/$ and 6500 gpm (421 kg/$, respectively. The cold-seawater 

temperature available for the condenser was determined by experiments at the STF to be 

Tcw = 6.1°C, which is lower than the original baseline value of 7°C. The nominal warmseawater temperature is 26"C, and it ranges between 24.5"C and 27.5"C during the year. 

The demonstrated performance of the heat exchangers in fresh water and seawater in the 

SERI OTEC laboratory and in the HMTSTA gives increased confidence in the following 

performance characteristics predicted for the NPPE: 

Evaporator thermal effectiveness = 

Tin - = 0.90 

Tin - Tsat , sw(Pevap) 
Overall direct-contact condenser thermal effectiveness = Tout - Tin = 0.86 . Tsat ,SW(~DCC) ' Tin 

The thermal effectiveness is the ratio of observed temperature difference in the seawater as it flows through the heat exchanger (the numerator), with respect to the total 

available temperature driving potential. TSat,sw(P) = Ts is the seawater temperature in 

equilibrium with the steam, at the respectlve operating pressures in the evaporator and 

at the entrance to the direct-contact condenser. Based on these values and on a heat 

balance of the system, as described in the system performance analysis (Section 2.3), the 

baseline steam flow was increased from the original baseline of 3.4 kg/s to a nominal 

condition of 3.5 kg/s. 

Concurrently, the steam temperature at the outlet of the diffuser that exhausts gas from 

the turbine was decreased from 12.4"C to Ts = 11.8"C. 

These are the given nominal operating conditions at which the heat exchangers were 

designed. 

5.2.2 Major Geometric and Operational Considerations 

Three key decisions were made at the inception of or during the course of the conceptual 

design effort which affect the design approach: 

1. The HMTEAJNPPE will use a dome configuration. 

2. A quantity equivalent to about 10% of the steam produced under nominal operation 

will be used to demonstrate the capability of producing desalinated water. 

3. Water production will be demonstrated after net power has been produced. 

These decisions, although they do not stem from design or performance considerations 

for the main heat exchangers themselves, do affect the design philosophy. 

The first decision was motivated primarily by turbine and structural considerations, as 

discussed in Section 2.4.3. This decision removed the design of the containment vessels 

from this task, because the entire HMTEAJNPPE system is integrated into one 

containment vessel, the dome. The structural requirements to provide a barrier to 

atmospheric pressure and to support the turbine are discussed in Chapter 3.0. Only the 

structural requirements to separate the evaporator chamber and the direct-contact 

condenser (DCC) stages remain in the domain of this task, to be coordinated with the 

structure team. 

The second and third decisions were made during discussions at the First Working Group 

Meeting (1989) regarding the use of the facility. Because the first priority is to produce 

net power, it was agreed that production of desalinated water would be a second priority 

that should not directly affect that goal. A quantity equivalent to 10% of the steam used 

for power production is sufficient to show proof of concept for desalinated water 

production and to significantly advance the technology. After net power is produced, the 

power requirement for water production will be assessed, as will the impact of using 

larger quantities of steam to produce desalinated water for larger systems. 
5.2.3 Secondary Heat Exchanger for Desalinated Water Production 

The type and location of the secondary heat exchanger to demonstrate desalinated water 

production was assessed early because of the potential impact on geometry, design, construction schedule, and operation of the main NPPE heat exchangers. 

Task members compiled a list of options and evaluation criteria (see Table 5-1). Options 

included designing a new surface condenser, moving existing HMTSTA equipment to the 

HMTEAINPPE location, and operating the water production system downstream of the 

NPPE system by using the seawater discharged from the NPPE into the existing HMTSTA 

equipment. 

A second consideration was whether the steam for water production should be taken from 

the evaporator before it flowed through the turbine or downstream of the turbine. The 

effect on performance of the latter was not considered in detail, but it results in a 

potential decrease of the turbine performance, depending on how the steam would be 

extracted. These penalties could result from pressure loss and flow separation near the 

steam removal point on the diffuser, as well as from steam maldistribution to the 

primary condensers, because of differences in pressure drop between the surface and 

direct-contact condensers and the difference in condensing temperature seen by the 

steam in the two cases. 

During the discussions, advantages and disadvantages were also identified for the location of the water-producing equipment. Locating the secondary condenser inside the 

dome may be more representative of larger OTEC systems and would have less visual 

impact. However, disadvantages include complex design of the main condenser, difficult 

water and steam routing, and difficult access for people installing or repairing the unit. 

For a unit outside the dome, the advantages are simple geometry, accessibility, easy 

water route, and operational flexibility. Disadvantages are the visibility of the external 

unit and the necessity of sealing one more pipe from the atmosphere. 

Based on the discussions, the working group members recommended that a new surface 

condenser unit be designed specifically for the application, that it be located outside the 

dome, and that steam be provided to the secondary heat exchanger from the evaporator 

using a dedicated spout. No provisions were included to route steam from the turbine 

exhaust diffuser to the surf ace condenser. The dedicated spout allows production of 

desalinated water with either the nominal warm-seawater flow rate (and lower than nominal steam flow to turbine) or with nominal steam flow rate (and higher seawater flow). 

This option provides for new technical development (a new design of the surface condenser) while minimizing the influence on the design of the primary heat exchangers 

resulting from geometry restrictions, performance, and development schedule of the 

secondary heat exchangers. It also eliminates any potential performance penalties on the 

turbine and main heat exchangers, whether or not fresh water is being produced. If 

difficulties arise developing or testing the new surface condenser unit, the other options 

in Table 5-1 could be reassessed and implemented if they necessitated only minimal 

changes in the design of the main heat exchangers. 

Documentation of discussions and ranking of the options is provided in SERI reference 

documents 2.2.3.2 and 2.2.3.4, two letters from ANL dated March 10 and 24, 1989, and 

one letter from PICHTR dated March 17, 1989. 
Table 5-1. Options for Secondary Condensers-Desalinated Water Production 

Options First Option Fallback Option Fallback Option 

Move existing equipment to NPPE Existing at HMTSTA 

Equipment Options Design new SC equipment Rosenblad Rosenblad DCC closed-loop 

Source of steam NPPE evaporator After turbine NPPE evaporator After turbine HMTSTA evap. HMTSTA evap. 

Capacity Design Design Design Cannot meet Design >0.35 kg/s 

design (0.25 max) 98% cond 

Schedule One year design One year design When applicable When applicable OE 

One year build One year build 

Pumps needed 

- CW Yes 

- WW NO 

- Condensate Yes 

- Closed loop No 

Yes 

No 

Yes 

No 

Yes 

No 

Yes 

No 

No 

No 

Yes 

No 

Yes 

Yes 

Yes 

No 

Yes 

Yes 

Yes 

Yes 

Gas removal 385 ICFM add. 3853 ICFM add. No impact Large impact Existing Existing 

8503 ICFM add. 

Estimated power use 

- Vacuum system 4 kW add. 4 kW add. 

- Pumps 7 kW 7 kW 6 kW 

Effect on turbine None Potential derat- None 

ing of turbine/ 

DCC 

Higher use 23 kW 23 kW 

6 kW 6 kW 7 kW 

Potential derat- None 

ing of turbine/ 

DCC 

Quality of water Select Select Change Change Filter 

material material storage storage water 

None 

Filter 

water 

Technical gain Data on Data on None 

plate-fin plate-f in 

None Staged OTEC Staged OTEC 

Aesthetics Visually Not visible Large Large Large equip- Large equipbest outside condenser condenser ment ment 

dome outside dome outside dome Large power use Large power 

use 

Cost estimate 

- Added equipment $50K $50K None None None $ 12K 

- Pipes Steam pipe Steam pipe 150 ft x 2 150 ft x 2 

- Pumps 2 each 2 each 2 each 1 each 3 each 3 each 

- Installation Install new SC Integrated Move and mount Move and mount None Add plates to 

$2K $2K closed-loop 

heat exchanger 

Steam flow = 0.35 kg/s. Drinking water quality required. Power used is to be accounted for in the NPPE balance. No interference with 

operation/construction schedule of NPPE main exchangers. 

DCC = direct-contact condenser; SC = surface condenser; NPPE = net power-producing experiment; HMTSTA = Heat- and Mass-Transfer Scoping Test 

Apparatus; OTEC = ocean thermal energy conversion 
5.2.4 Predeaerators 

Predeaeration of the warm and cold seawater has been proposed to reduce parasitic 

losses in OC-OTEC systems. The net power gain stems from a reduction of pumping 

power required in the vacuum exhaust system; gas released in the predeaerator would be 

at a higher pressure than the condenser exhaust pressure, requiring less energy for 

recompression up to ambient conditions. A secondary gain would be obtained by slightly 

increased effectiveness in the condenser because of lower noncondensable inlet gas mass 

fraction and no additional release of noncondensable gases in the stages of the DCC. 

This would result in a slightly smaller steam flow to the precooler. 

The parasitic power consumed in an OC-OTEC plant for compression of all noncondensable gases released during the process has been estimated to be less than 30% of the 

overall parasitic losses of the plant (Block and Valenzuela 1985). Taking hydraulic and 

vacuum system losses into account, the gain potential for predeaeration is on the order 

of 10% to 15% of the parasitic losses of the plant. This amounts to between 3% and 4.5% 

of gross power for a large power-producing system and is clearly a second-order gain. 

Analysis indicates that warm seawater predeaeration in the NPPE could contribute a 

maximum potential gain in net power of 12 kWe x F, linear in F where F is the fraction 

of the noncondensable gases in the seawater that are liberated in the predeaerator. This 

assumes that no hydraulic losses are incurred in the predeaeration process and that 

predeaeration occurs at 11.3 kPa (the inlet pressure of the third stage of the NPPE 

vacuum system). Predictions for both warm- and cold-seawater predeaeration show that 

the maximum potential gain is 20 kWe x F, where F is the weighted average fraction that 

is released from both seawater streams. 

Warm seawater predeaeration tests conducted at the HMTSTA during Phase I (Zangrando 

et al. 1990) at liquid loadings equivalent to the NPPE evaporator loading show F = 0.47, 

giving a maximum potential gain of 12 x 0.47 = 5.6 kWe. When hydraulic losses are taken 

into account, the net power gain is less than 4 kWe. 

Preliminary tests conducted recently at the HMTSTA indicate that predeaeration levels 

of 80% to 90% may be attainable. If these values could be obtained for higher liquid 

loadings than tested, and for cold seawater as well as warm seawater, then the maximum 

potential gain from predeaeration of both streams may be as high as 20 x 0.85 = 17 kW,, 

minus hydraulic losses. This is about 10% of the NPPE overall parasitic losses, In 

agreement with the estimates made for larger plants. 

There remains considerable uncertainty regarding the projected effectiveness of a 

predeaerator; this results in difficulties in apportioning the noncondensable gas flow to 

the various stages of the vacuum exhaust system. Should the capacities of the vacuum 

system stages be mismatched, significant power penalties (i.e., more than what could 

potentially be gained using a predeaerator) occur. Although a maximum potential gain of 

17 kWe from predeaeration is significant, this estimate uses an optimistic projection of 

the predeaerator performance. The more conservative estimate of 4 kW potential 

power saving does not warrant the complications and operational risk 07 including 

predeaeration for the NPPE. 

In addition to the technical uncertainty and given the existing knowledge of predeaerator 

performance and operation, serious design complications would be introduced by inclusion 

of predeaerators, including the following: 
a Probable need for separate, large predeaeration vessels. The low liquid loadings used 

to obtain high release rates in the HMTSTA require predeaerators with about twice the 

planform area of the NPPE evaporator. 

Design of the vacuum system to accommodate varied mass flow through the stages. 

This could seriously affect start-up. 

a Additional uncertainty in estimating how the NPPE vacuum pump should be sized. 

Currently, it is sized for 100% release at the condenser operating pressure. 

a More cumbersome water routing to bring the seawater to and from the predeaerators. 

The expected amount of increase in net power is small and does not justify the increased 

risk, complexity, and cost of hardware. Therefore, no predeaerators (for warm or cold 

seawater) will be used in the NPPE. 

Calculations of net power savings for the NPPE are documented in SERI reference document 2.2.3.9. 

5.3 Conceptual Design of Heat Exchanger Components 

The main heat exchangers consist of a spout evaporator and two DCC stages. The 

evaporator consists of a cylindrical annulus around the two DCC stages. The DCC stages 

are coaxial; the first stage is cocurrent, the second countercurrent. The countercurrent 

stage is at the center. The layout is shown in Figure 2-6. 

Because the dome configuration has been selected, the heat exchangers are contained in 

a single vacuum enclosure. Structural requirements for this enclosure are discussed in 

Chapter 3.0. The heat exchangers are designed on the basis of results obtained during 

the HMTSTA tests with seawater. These tests were conducted over a broad test matrix 

that envelops NPPE nominal operating conditions. 

Detailed computer models of heat and mass transfer in DCCs using structured packing 

have been verified extensively using fresh-water test results obtained at SERI 

(Bharathan, Parsons, and Althof 1988). These models have been modified for operation 

with seawater and have been also verified using HMTSTA test results (Zangrando et al. 

1990) These results have been used in an overall system model to predict the effect of 

heat exchanger performance parameters on the NPPE system. 

Detailed computer models have been used at NPPE operating conditions for the conceptual design of the condenser and to predict the nominal operating parameters. They will 

also be used during the preliminary design to fine-tune the geometry and operating 

conditions of the two stages of the DCC. 

The remaining components are the mist eliminator and the water distributors. The turbine diffuser will act as steam distributor to the first stage of the DCC. The small 

transition in flow area from the base of the diffuser to the first stage will not affect the 

condition of the steam; it is much smaller than that tested at the HMTSTA, where no 

measurable effect was observed. Neither warm- nor cold-water predeaeration is planned 

for the NPPE, as is discussed in Section 5.2.4. 

The combined requirements of planform area for the DCC, evaporator, and dividing walls 

between these heat exchangers define the required inner diameter of the vacuum vessel, 

6.4 m (21 ft). 
5.3.1 Evaporator 

The proposed conceptual design of the evaporator is shown in Figures 5-2a and 5-2b. The 

evaporator consists of a cylindrical annulus between the dome wall and an internal partition that separates the evaporator from the DCC. Its location in relation to the DCC is 

selected because of the radial-inflow turbine geometry. A horizontal distribution 

manifold provides warm seawater to multiple vertical spouts. The warm seawater 

exiting the spouts flashes and produces steam that rises up toward the inlet to the 

turbine. The remaining seawater drains into the space around the distribution pipes and 

exits the drainage plenum through a discharge pipe. 

The evaporator receives seawater at nominally 26°C and produces steam at 21.8"C. The 

nominal evaporator thermal effectiveness is 0.9; therefore, the seawater exits at 

22.5"C. The flow rate through the evaporator is 620 kg/s. Control of warm seawater 

flow to and from the evaporator occurs outside the dome, as is discussed in Chapter 4.0. 

The external flow controls are used to adjust the surface level of the drain pool with 

respect to the tops of the spouts. 

The main considerations in evaporator design are liquid loading, active spout height, and 

homogeneous distribution of water to the spouts. Liquid loading is defined as the mass of 

warm seawater per second distributed over a unit planform area of the evaporator. The 

active spout height is the distance between the top of the spout pipe and the surface of 

the drain pool. These parameters affect the quantity of steam produced, overall thermal 

effectiveness, and the homogeneity of steam distribution to the turbine. 

5.3.1.1 Evaporator Planform Area and Loading 

The evaporator's area requirement is governed by considerations of liquid loading and 

accessibility. In the HMTEAJNPPE dome, accessibility is the more stringent constraint. 

The evaporator is an annulus that must be wide enough that a person can comfortably 

enter and work in it. 

Tests at the HMTSTA have consistently shown evaporator effectiveness at 9r above 0.9, 

for single as well as multiple spouts, at liquid loadings up to about 50 kg/m s and active 

spout heights of 0.5 m or more (Figure 5-3). These results agree well with results 

obtained by French scientists (see discussion in Zangrando et al. 1990 and SERI reference 

documents 2.2.3.3 and 2.3.3.5). For the same steam loading and water loading the same 

performance is obtained, regardless of spout size or number of spouts. Both French data 

sets fall within the error bar of the results obtained at the HMTSTA with seawater. T9s 

confirms that high thermal performance is attainable for liquid loading around 50 kg/m s 

from four different test facilities, using both fresh water and seawater. 

A liquid loading of 50 kgpLs at nominal warm water flow of 620 kg/s requires an 

evaporator area of 12.4 m . During the first working meeting it was agreed that the 

design would include a provision to shield each spout at a diameter near 1 m (the 

diameter of the HMTSTA is 3.5 ft), so that each spout could operate in an environment 

similar to that of the HMTSTA, if needed. Given the inner diameter of the dome wall, 

6.4 m (21.0 ft), and requiring a width of about $m for the annulus, the area of the I HMTEA PPE evaporator will be about 17.7 m , for an overall liquid loading of 

35 kg/m P s. This liquid loading is quite low, and it should result in a thermal 

effectiveness higher than 0.9. 

The planform area of the evaporator can be adjusted around the value of 17.7 mL if construction considerations can benefit from small changes in this area. The width of the 


m 

hnnulus should not be reduced be- @ low 1 m-the usual width of corridors--for accessibility and ease 

of manifold installation. The 

outer diameter of the evaporator 

chamber is fixed by the inner 

diameter of the dome. 

5.3.1.2 Evaporator Spouts 

Twelve spouts 0.24 m (10 in.) in 

diameter satisfy the geometric 

and loading constraints. Because 

spout diameter has been shown 

not to affect the performance of 

the evaporator (Zangrando et al. 

1990), the largest spouts tested at 

the HMTSTA have been selected 

for the NPPE, to reduce the hyFigure 5-3. Influence of liquid loading on draulic losses for a given liquid 

seawater spout evaporator loading. An additional spout 

dedicated to the surface 

condenser may be shielded from 

the other spouts. The 12 main spouts are distributed symmetrically around the annulus 

for even distribution of steam. They are placed evenly along the distribution pipes 

(except along the axis of the main water supply pipe), where they provide sufficient 

space in the enclosure for easy access from below as well as for side access into the DCC 

chamber. The configuration provides for balanced flow to the two arms of the warmwater distributor, whether or not the spout dedicated to the surface condenser is active. 

The minimum active spout height is 0.5 m. The water distribution pipes may be installed 

farther below the spouts to provide flexibility in drain-pool level and to accommodate 

tidal variations, if needed. If the level of the pool's surface needs to be lowered slightly 

to accommodate unforeseen pump requirements, the spouts can be trimmed 

appropriately. 

The spout dedicated to the surface condenser is mounted directly above the single vertical supply pipe. It is of the same diarrteter as ihe other spouts, but it is fitted with a 

water control valve to regulate the seawater flow for steam production to the surface 

condenser. 

5.3.1.3 Evaporator Water Supply and Distribution 

The water distributor requires a single penetration through the dome wall. A pipe of 

0.76-m (30-in.) nominal diameter enters horizontally and turns vertically inside the 

evaporator discharge plenum, as shown in Figures 5-2a and 5-2b. This pipe is connected 

to the center of a C-shaped horizontal distribution pipe. The nominal diameter of the two 

branches is 0.61 m (24 in.). The vertical spouts are mounted at regular intervals on these 

two branches. The water distributor is supported on a wide-mesh grate mounted on the 

separating wall and the dome wall. A pipe distributor rather than a supply chamber is 

chosen for ease of construction and drainage of the spent seawater. Velocities in the 

NPPE spouts are in the range tested at the HMTSTA. 
Penetrations through the dome walls should be limited to reduce the number of possible 

sites of leakage. The selection of pipe sizes is such that the area of the two branches is 

approximately equal to the area of the single vertical supply pipe, for even distribution I between the two branches. The vertical pipe size is also selected to facilitate 

installation in the narrow evaporator plenum. Some balancing of flow rates among spouts 

can be obtained by trimming the heights of the spouts on site, if necessary. 

Other configurations were evaluated but discarded, because analysis of the configuration 

presented here showed satisfactory results for hydraulic losses and flow homogeneity (see 

Sections 5.3.1.4 and 5.3.4). A larger pipe for the two branches of the evaporator water 

distributor was considered but eliminated as unnecessarily large. In addition, increasing 

- the diameter of the branches would restrict drainage from the evaporator because of the 

narrowness of its chamber. 

Tests at the HMTSTA (and at two French research facilities) operating with multiple 

spouts have shown no spout instabilities, even in the presence of two-phase flow. 

Hydraulic losses in the spouts have been measured in the HMTSTA for spouts of the same 

size as those proposed for the NPPE. The differences in design of the water distributor 

proposed for the NPPE, compared to that tested at the HMTSTA, are not sufficient to 

result in substantially different behavior, given the low hydraulic losses. Therefore, most 

task members consider the hydraulic design of the water distributor to be 

straightforward. 

5.3.1.4 Flow Variation in the Evaporator Distributor 

PICHTR performed a detailed analysis of the possible variation in flow among the spouts, 

assuming a constant loss coefficient. Results show that the maximum variation in flow 

between spouts is within 4%, below the target value of about 5%. Details of the geometry and calculations performed are provided in a PICHTR letter dated November 3, 

I 1989. 

5.3.1.5 Evaporator Discharge Plenum 

The discharge plenum allows for ease of collection of the spent warm seawater and for 

possible separation of foam or bubbles before the water is discharged to the drainpipe. It 

also provides a way of easily controlling surface level of the discharge pool. The plenum 

fills up from empty to the nominal level in less than two minutes at nominal flow. 

The nominal level, governed primarily by seawater systems constraints, is 4.1 m (13.5 ft) 

above grade. The constraints from the heat exchangers' point of view are to maintain 

the required active spout height; to provide sufficient room for installation of the water 

supply pipe, distributor, and instrumentation; and to allow for easy access to the 

evaporator chamber. 

The spent seawater discharges through a pipe near grade that is 0.91 rn (36 in.) in diameter. Access to the evaporator chamber is provided by a hatch through the dome wall 

near grade level and through the grating that supports the water distributor (see 

Figure 5-Za), at a position opposite the dedicated spout. 

The drain plenum has two walls: the dome wall on the outside, and the cylindrical wall 

between the evaporator and the DCC. This inner wall must support the load of the 

diffuser from the turbine as well as the hydrostatic load resulting from the difference 

between pool surface levels in the evaporator and DCC during operation. Additional 

strength will be provided for this inner wall because, at start-up, one of the drain 
plenums may be filled before the other. The inner wall also will have three penetrations 

for the DCC's supply, discharge, and vent pipes. The height of this wall is determined by 

the highest level the warm seawater is expected to reach. A lightweight deflecting 

surface placed above the top of this wall ensures that seawater splashing from the spouts 

will not reach the turbine diffuser walls. 

Water leakage between the drain pools should be prevented, but it is not very significant 

as long as discharge seawater temperatures are monitored upstream of the leak. Leakage of warm seawater into the DCC stages should be prevented, because it would 

seriously affect the performance of the DCC. 

5.3.2 Mist Eliminator 

The geometry of the evaporator chamber in the HMTEAINPPE dome configuration 

provides for very low steam velocities, about 6.2 m/s, at the flange of the turbine. The 

distance between the tops of the evaporator spouts and the turbine inlet is also great, 

about 2.7 m (9 ft), so that substantial gravitational settling of entrained droplets will 

occur. In addition, the evaporator liquid loading is low enough that only a small quantity 

of entrained droplets is introduced into the steam at the level of the evaporator spout. 

On the basis of data collected at the HMTSTA, a mist eliminator may not be necessary 

for the NPPE. However, its use is planned because it can be installed readily in the 

NPPE dome, because of its low cost, and because the dynamic losses are low. The unit 

may be removed after the HMTEA tests are completed if benefits can be identified. 

The mist eliminator tested at the HMTSTA is a Flexi-Chevron Type VIII, made of 304 

stainless steel with 25-mm (1-in.) spacing between the chevron passages. The NPPE mist 

eliminator will be of the same type but made of plastic, which is lighter and less costly. 

Although it has 19-mm (314-in.) passages, the dynamic loss for the NPPE remains small 

because of the low steam velocity. 

The mist eliminator will be installed horizontally between the dome wall and the turbine 

exha t diffuser at a height near the turbine flange. The area at this location is about UrS 28 m . Positioning is not critical, but it should be close to the turbine inlet and should be 

selected for ease of installation and support. 

The pressure loss coefficient for the plastic chevron mist eliminator is 10 (Zangrando 

et al. 1990). For a steam velocity of 6.5 m/s at the height of the mist eliminator, the 

total pressure drop is 4 Pa. 

5.3.3 Direct-Contact Condenser (DCC) 

The proposed design of the DCC for the NPPE is shown in Figure 5-4. Two levels of the 

DCC plan view are shown in Figure 5-2a, along with other components. In the first stage 

of the condenser, the seawater and the steam flow cocurrently. This stage condenses 

most of the steam exhausted from the turbine. In contrast, the second stage is 

countercurrent, with seawater flowing down and gas flowing up. This stage condenses 

the remainder of the steam down to near-equilibrium conditions, and it concentrates the 

noncondensable gases. The second stage is smaller in diameter than the first stage. It is 

located at the center of the vessel so that the gas to be exhausted can be collected into a 

pipe easily and with small pressure drop and routed to the vacuum system. The first 

stage is an annulus around the second stage. Because the first stage is longer in 

diameter, the annulus is wide enough to have minimal wall effects. Steam from the 

turbine is routed to the first stage by a conical diffuser, which acts as a steam 

distributor. 

Cold seawater is distributed on top of both stages and spreads through the structured 

packing in each stage. The spent water is collected below the two stages in a drain 

plenum common to both and is discharged through a pipe at the bottom. 

The area below the structured packing of the two stages, down to the surface of the 

drain pool, is open to allow passage of the uncondensed steam from the first to the 

second stage. The flow area between the stages is approximately the same as the 

planform area of the second stage to minimize gas pressure losses as the uncondensed 

steam turns around. 

The design is virtually the same as that tested in the HMTSTA during the Phase I1 coaxial 

tests, except the DCC is proportionally bigger. 

The DCC receives cold seawater at nominally 6.1°C. The total flow to the condenser is 

421 kg/s, split appropriately between the first and second stages. The condenser's overall 

nominal thermal effectiveness is 0.86; therefore the spent seawater exits at an average 

temperature of 11.3"C. At nominal conditions, the condenser exhausts gas at a vent 

ratio of 0.9. The vent ratio compares the ideal (smallest) volumetric flow of steam that 

must be exhausted with the noncondensable gases with the actual volumetric flow rate 

removed and is defined as 

'i, 1 - CAP - P 

v = - sat(Tso) .- Twi 

'i,l Psat(Twi ,2) Tso 

A vent ratio of 0.9 implies that the volume of steam exhausted is about 10% larger than 

the theoretical minimum. 

The main considerations for design of the DCC are steam loading to the first stage, stage 

area ratio, liquid loadings, packing and freefall height of both stages, interstage gas flow 

area, and evenness of water distribution to the stages. The steam loading is the mass of 

steam per second that enters a unit planform area of the condenser stage. The stage 

area ratio compares the planform areas of the first and second stages. The liquid loading 

is the mass of seawater per second distributed over a unit planform area of condenser 

stage. The packing height is the physical height of the structured packing; the freefall 

height includes the heights above and below the packing at which seawater is in contact 

with the steam. These parameters affect overall thermal performance, approach to 

exhaust equilibrium conditions of the steam, and overall pressure losses. 

Computer model verification of predicted NPPE performance and a preliminary 

evaluation of the area ratio and water loading of the DCC stages were obtained with the 

detailed SERI computer models for each stage (Bharathan, Parsons, and Althof 1988), 

which have been verified with HMTSTA data for operation with seawater (Zangrando 

et al. 1990). 

5.3.3.1 DCC Planform Area and Required Loadings 

On the basis of performance during th HMTSTA tests, the steam loading to the first 5 stage at nominal conditions is 0.4 kg/m s2 For the nominal steam flow of 3.5 kg/s, this 

fixes the area of the first stage at 8.75 m . A preliminary evaluation of the tradeoffs in 

overall condenser performance with respect to stage area ratio and water loading of the 

stages indicates that the area ratio of 3 used in the HA$T'STA gives good performance. 

Therefore, the nominal area of the second stage is 2.92 m . At these conditi s, best use OT of the water is attained with a liquid loading to the first stage of 38 kg/m s. Given a 
fixed quantity of cold w ter available, the above parameters fix the liquid loading of the 9 second stage at 30 kg/m s. The resulting nominal seawater flow rate to the first stage is 

333 kg/s, and to the second stage it is 88 kg/s. 

The outer diameter of the DCC depends slightly on the sizes of pipes at the center of the 

structure, the area of the two stages, and the thickness of the partition that separates 

them. This partition must provide an impermeable barrier to steam and water flows in 

the two stages. It also may need to support the internal cone of the diffuser, if it is not 

supported by the outer flange of the diffuser or the outer wall of the DCC; guidance 

from the turbine diffuser's designer will be obtained during the preliminary design. This 

partition also may support the water tray, if the tray is not installed onto the packing 

itself. Depending on decisions made during the preliminary design, this partition can be 

very thin. Sufficient space has been included in the dome's inner diameter for the 

support walls. 

Figure 5-5 shows the predicted overall thermal effectiveness as a function of water 

loading of the first stage and stage area ratio (indicated by the number used as symbol). 

These calculations were made for a fixed nominal water flow rate, constant steam 2 loading of 0.4 kg/m s, and noncondensable-gas mass fraction of 0.3% at the inlet to the 

first stage. Results for two heights are shown in the figure: the minimum height of 1 

structured packing and the total freefall height of each stage, as discussed below. Some 

flexibility exists in the actual stage area ratio and related liquid loading if structural 

considerations indicate a benefit from some small adjustment to the stages' horizontal 

dimensions. The effects of such modifications need to be modeled during the preliminary 

design before the modifications are made. 

5.3.3.2 Structured Packing and Condenser Heights 

The same packing used in the HMTSTA is planned for the NPPE--a MunJers CF-25060 

structured packing with effective surface area per unit volume of 98 m /m3. Packing 

heights are 0.61 m (2 ft) in the cocurrent stage and 0.91 m (3 ft) in the countercurrent 

stage. The packing used at the HMTSTA is composed of sections 0.15 m high that are 

rotated in the horizontal plane with respect to the neighboring sections to obtain a different orientation of the flow channels. If detailed modeling during the preliminary 

design shows that a measurable decrease in steam temperature can be obtained at the 

exit of the countercurrent stage by small modifications of the packing geometry, a single 

stack of slightly different packing may be substituted at the top of the second stage. 

This would not affect the stage's overall height. 

The total freefall height will be virtually the same for both stages, 1.5 m (5 ft). The 

height of the interstage gas passage is controlled by the area selected for the countercurrent stage. At nominal conditions it is 0.47 m (1.5 ft) high. 

Computer model predictions for these heights give a possible overall thermal 

effectiveness of 0.89, an overall pressure drop of 83 Pa, and a steam saturation 

temperature at the exit of the countercurrent stage of 6.0°C (the theoretical limit for 

operation with seawater is 5.82OC). However, the original conservative nominal value of 

0.86 for overall thermal effectiveness of the DCC should be used in the conceptual 

design calculations, because the predicted increase is within the error bars for model and 

seawater data. 

For overall pressure drop, comparison of the model predictions to all coaxial data 

obtained in the HMTSTA yielded a standard deviation of k16 Pa for the cocurrent stage 

and +19 Pa for the countercurrent stage. These are very small pressure variations but - 
5.3.3.3 Water Distributors for the DCC 

they amount to a proportionally large 

The water distributors should be simply designed on the basis of configurations tested at 

the HMTSTA. They should occupy a small area and should not interfere with the gas flow 

to the first stage or with the gas flow between the stages. Because most of the area 

above the countercurrent condenser is unused except to vent the gas mixture, it is 

profitable to place the water distributors in this area. Dry water pipes in the cocurrent 

stage should contribute only a small amount of steam blockage (on the order of 10% of 

the area as in the HMTSTA coaxial configuration), so that the steam pressure drop 

around the pipes remains small. However, if the surface of the pipes is wetted, the proportional area of the pipes can be larger. In this case, the steam condenses onto the cold 

seawater that emanates from the pipes and this acts as suction to keep the steam flow 

attached (e.g., Schlichting 1968). 

error in pressure drop predictions 0.900 

(Zangrando et al. 1990). The gas 

pressure drop data from the HMTSTA 

must be sorted, during the prelim- (I] 0.895 

inary design phase, to isolate those (I] a, 

c 

data near the NPPE nominal .- - 2 

conditions for which the drift in the h: 0.890 

w pressure drop sensors had been - 

closely monitored. Until this analysis 2 a, 

is completed, the entire uncertainty 6 0.885 

between model and data will be 

assigned to the model predictions. 

This conservative approach results in 0.880 

The pro osed configuration is to introduce the cold water with a single supply pipe of 

0.61-m P 24-in.) nominal diameter that penetrates through the dome wall and through the 

inner wall separating the evaporator from the DCC, as shown in Figures 5-2a and 5-3. In 

the condenser plenum, this pipe turns vertically at the center of the structure and carries 

all the cold seawater to just above the second stage, where it is distributed to both 

stages. Control of water flows is discussed in the next section. 

The reason for proposing a combined distribution system from the top of the second stage 

is its simplicity. It takes advantage of the unused space above the second stage, the low 

pressure available at the top of the second stage, and the freefall height available above 

the first stage. In this manner, all the cold water is brought in through a single pipe, and 

all the water is exposed to the condenser exit pressure. The cold water will be deaerated, slightly improving the DCC's performance. 

a predicted overall pressure drop of 30 32 34 36 38 40 

83+35 Pa for the NPPE. Therefore, ~irst-stage liquid loading (kg/m2s) 

until a more realistic prediction is 

made, the nominal overall pressure Figure 5-5. Model parametric runs for 

drop will be taken as <I20 Pa. The NPPE, packing and freefall heights 

gas pressure drop affects prediction 

of the vent performance for the DCC. The net effect of the uncertainty in pressure drop . results in a preliminary specification of steam flow to the precooler that is larger than 

expected during operation, for conservative design. 

# 

0 

8 

4 

*>-A - 

-4-4-4 

- I( /- 

Aspect ratio 

A 3 ,dA\ m2 

- .4 F --- Packing height / 

Freefall height 

I 

I 
Commercial water distributors for cooling towers and packed columns are available that 

use design concepts similar to those considered here. However, a custom design is 

preferred to provide greater flexibility, and it may cost less. 

The water distribution system to the second stage is well defined and can be scaled up 

from the geometry successfully used in the HMTSTA. This consists of a perforated drip 

tray with symmetrically distributed gas standpipes. The holes have 5-mm (13164-in.) 

diameters and are evenly spaced in a square pattern at a hole concentration of 6.5%. 

The tray is installed directly above the second stage and has the same planform area as 

this stage. It can be mounted roughly 15 cm (6 in.) above it. At the liquid loadings 

required for the NPPE second stage, the water level in a tray configured like the one 

tested at the HMTSTA is about 40 mm (1.5 in.). If greater levels are needed for the 

first-stage distributor, the number of holes in the tray can be reduced accordingly. 

Nominal height from above the packing to the surface in the water tray is 0.3 m (1 ft), to 

be apportioned as needed. 

Water distribution to the first stage is not fully defined, but it will be similar to that 

used in the HMTSTA coaxial configuration tests, in which the vertical distribution pipes 

were supplied from a ring distributor installed just below the discharge pool level. This 

location was selected at the HMTSTA only to ease installation. In the NPPE, the recommended alternative is to build the distributor as an integral part of the second-stage 

water distributor. Equally spaced pipes protruding radially outward from the edge of the 

water tray will distribute the cold seawater over the first stage. Thirty-six pipes would 

sufficiently distribute all the water for the cocurrent stage, maintaining about the same 

ratio of coverage that was tested at the HMTSTA. The radial distribution pipes will be 

perforated along their length to offer a wetted surface to the incoming steam, resulting 

in negligible steam blockage. Details of pipe size, water level in the tray, perforation 

size and number, and relative elevations will be determined during the preliminary 

design. 

Other alternatives are under consideration and a recommendation will be made during 

the preliminary design (see SERI reference number 2.2.3.8). One alternative is to 

separate the flow to the two stages before the cold seawater is introduced into the 

DCC. In this case, the first-stage water distributor can be installed in the space above 

the second stage or under the first stage, submerged in the discharge pool. Installation in 

the evaporator plenum has been considered but leakage is a serious concern. 

5.3.3.4 Control of Water Flow 

The overall flow rate to the DCC will be controlled outside the vessel on the main supply 

line from the pump station. Inside the vessel, therefore, it is only necessary to fine-tune 

the relative amounts of water to the stages. During the HMTEA tests, the design is for a 

variation of -50% to +lo% of nominal flow, but tests will be conducted at only three or 

four overall flows. At each setting, the split between the two stages may be changed by 

about ?lo%. During the NPPE tests the overall flow will vary by no more than 210%. 

Therefore, HMTEA test requirements should not complicate the design of the seawater 

systems if satisfactory control of the water flows can be obtained with a single supply 

pipe for both stages. 

Note that during the NPPE tests, it is not necessary to directly measure the flow rate to 

each stage. Flows can be adjusted by observing the water temperature difference across 

each stage. During the HMTEA tests, the flow rates can be measured with techniques 

other than flowmeters in the supply lines. Therefore, from an operational point of view, 

separate pipes are not required. 
Water flow to each stage can be measured and controlled in several ways. The simplest 

method in the proposed configuration is to restrict the entrance area of the first-stage 

distribution pipes from within the water distribution tray, and to monitor surface level of 

water in the distribution tray and velocity in the radial distribution pipes. This system 

can be calibrated before tests begin when the countercurrent stage is installed. Details 

of the control and measurement strategies still need to be defined. Restriction of the 

second-stage flow has also been considered. 

In the recommended design, the flow to the two stages is controlled by the surface level 

of water in the tray above the second stage and by the hydraulic losses through the tray's 

bottom and the first-stage radial distribution pipes. These sections can be designed to 

obtain the desired nominal flow split out of the common tray by selecting proper sizes 

for the respective flow passages. Trimming of the water split requires a device that 

slightly modifies the hydraulic losses into one of the two distributor sections by varying 

the flow-path area. The technique selected must ensure that the control is applied 

evenly across all openings of either distributor, without significant individual hysteresis, 

so that even flow is maintained to all sections of each stage for a particular flow rate. 

Several alternatives are under consideration. Flow to the first-stage radial pipes could 

be controlled with commercially available pinch valves, short pipe spools with internal 

flexible sleeves that expand and contract pneumatically from a common air source, or 

with other inflatable obstructions. V-shaped open distribution channels are also under 

consideration. A cylindrical ring inside the water tray could be rotated or lowered to 

block part of the entrance for the radial pipes. Control of flow to the second stage may 

consist of a plate attached below the water tray, with the same hole pattern as at the 

base of the tray, mounted on a rotating axis and moved by a gear mechanism. The plate's 

position would control the available flow area. Recommendations on these alternatives 

will be made early in the preliminary design. 

The use of activated butterfly or ball valves on each radial pipe supplying the first stage 

was discussed, but that appears to be beyond the level of control required. 

5.3.3.5 Discharge Plenum of the Direct-Contact Condenser 

The spent seawater from the two stages drains into a common plenum below them and 

discharges into a common pipe which penetrates two walls: the wall that separates the 

warm from the cold-water discharge plenums and the dome wall. The diameter of this 

discharge pipe is planned at 0.76 m (30 in.) to facilitate discharge of possibly foamy 

seawater, as was observed in the HMTSTA. 

The nominal pool height, 2.1 m (7 ft), is governed primarily by seawater systems and 

structure constraints. The constraints from the heat exchangers' point of view are to 

maintain the required freefall height, to provide sufficient room for installation of the 

seawater supply pipe and exhaust steam pipe and for instrumentation, and to allow for 

easy access to the DCC chamber. Such access will be through a hatch in the wall that 

separates the evaporator from the condenser drain pools. Access to the top of the 

condenser may be through the central water supply pipe or perhaps through the turbine 

diffuser walls from the evaporator chamber. In the second case, care must be taken to 

ensure that steam does not leak through the diffuser openings. Appropriate selection will 

be made during the preliminary design. 

A few supports may be required in the plenum for the wide-mesh grate that supports the 

packing, the partition between stages, and the water distributor. 
5.3.3.6 Gas Vent Pipe 

The second pipe shown in Figures 5-2a and 5-4 near the center of the countercurrent 

stage removes the exhaust gases to the vacuum system. Its only requirements are that it 

be high enough to ensure that water cannot be spilled into the vacuum system and large 

enough to contribute only a very small gas pressure drop between the DCC and the 

vacuum system. Nominal diameter for this pipe is 0.61 m (24 in.). 

5.3.4 Hydraulic Losses of Main Heat Exchangers 

The upper estimates for the overall hydraulic losses to be incurred in the evaporator and 

DCC described above are as follows: 

Static Dynamic Total 

Evaporator 0.5 m 0.35 - 0.5 m 1.0 m 

Direct-Contact 

Condenser 1.7 m (5.5 ft) 0.35 - 0.5 m 2.2 m 

These estimates are conservative, and values have been rounded off. The amount of 

these hydraulic losses attributable to discharge pipes is very small, about 0.05 m for each 

of the discharge pipes. This is smaller than the range of uncertainty in the dynamic 

losses on the supply sides. However, the overall losses in the heat exchangers are so 

small compared to the overall frictional losses in the other piping that pump selection 

will not be affected by this uncertainty. 

The static losses take into account the height of the water column from the discharge 

point down to the surface of the drain pool. The estimates assume that four separate 

pipes are available outside the dome, two for supply and two for discharge of the warm 

and cold seawater. The estimates incorporate only the losses from just outside the dome 

through the heat exchangers. The last column assumes worst-case conditions in the 

pipes, namely some regions of two-phase flow at the lower pressures, as was observed in 

the HMTSTA. 

The range in evaporator dynamic losses depends on the specific location and size of 

reducers, the fraction of two-phase flow to be encountered in the distribution pipe and 

spouts, and whether or not the spout base is flared. Results at the HMTSTA have shown 

hydraulic loss coefficients for the free-standing vertical spout of about 1.9 with bellmouth entrance and about 3 without. It is believed that two-phase flow effects occur 

only in the upper region of the spout. Therefore, the frictional loss in a large part of the 

vertical spout may result only from single-flow pipe friction. The HMTSTA spout 

frictional losses were taken into account in the estimates. 

Dynamic losses for the evaporator were calculated independently by PICHTR to be 0.4 m 

for the supply side and 0.04 m for the discharge side, for a total of 0.44 m, as compared 

to SERI's worst-case estimate of 0.5 m. These are in excellent agreement. The minimum losses for the DCC distributor, assuming single-phase flow, were calculated by 

PICHTR to be 0.14 m for supply and 0.04 m for discharge, for a total of 0.18 m, compared to SERI's calculated minimum of 0.35 m. For two-phase flow conditions near i.le 

point of discharge of the cold water, PICHTR estimates dynamic losses of 0.19 m plus 
0.04 m for a total of 0.23 m. The SERI estimate is 0.39 m. These discrepancies appear 

to stem primarily from slightly different assumptions. Because some details are still to 

be worked out in the piping to the DCC, a conservative upper limit of 0.5 m is 

maintained for the conceptual design. Documentation of the detailed calculations can be 

found in a PICHTR letter dated November 3, 1989. 

5.3.5 Surface Condenser for Desalinated Water Production 

The proposed secondary condenser for production of desalinated water is an extrudedchannel surface condenser with finned surface on the steam side. The unit is made of 

brazed aluminum and contains three water passes with cross-flow of steam. The number 

of fins per unit length is different for each pass, as shown in Figure 5-6. Figure 5-7 

shows the unit's dimensions and the required piping connections. This unit is designed on 

the basis of computer models that have been verified at the HMTSTA for a dimple-plate 

configuration. The design of this unit is beyond the conceptual stage, and all details are 

provided in a separate report (Panchal and Genens 1989). 

Cold seawater is supplied from a tee in the main supply line to the DCC. Spent seawater 

is discharged into the main drain pipes. The pipe size for seawater supply and discharge 

is nominally 0.3 m (12 in.); the condensate is sent to storage tanks through a 0.1-m (4-in.) 

pipe. Separate pumps are used in the loops for seawater and desalinated water. 

Steam is supplied from the dedicated spout. This spout is similar to the other spouts, but 

there is a water control valve at the base of the spout and flexible, light shielding around 

it to isolate it from the evaporator enclosure. This shielding is connected to a 0.61-m 

(24-in.) penetration in the vessel wall above the evaporator and provides steam to the 

surface condenser through a valve such as the butterfly valve in the HMTSTA. The 

uncondensed steam and noncondensable gases are vented to the precooler of the vacuum 

system through a 0.2-m (8-in.) pipe with a control valve. Capability to control the steam 

mass flow rate and the steam pressure separately increases the testing range for the 

unit. As in the HMTSTA, steam to the surface condenser can be controlled by adjusting 

warm-water flow and pressure drop through the steam valve. If flexible shielding were 

removed for some tests, evaporator chamber steam would flow to the surface 

condenser. However, independent control of steam flow rate and pressure would not be 

possible in this operational mode. 

At nominal conditions, the surface condenser receives 0.35 kg/s of steam and 10% of the 

cold-seawater flow. The unit is predicted to condense 99% of the steam at these conditions. 

5.4 Design Performance and Uncertainties 

Design performance has been discussed in the text and is summarized in Table 5-2 for the 

main heat exchangers. The values listed are the conservative estimates for nominal conditions. The values in parentheses are the potential gains that are expected for the 

NPPE design. 

The largest uncertainty is in the expected pressure drop of the DCC. This value is small 

and difficult to measure. Its effect is seen in the vent ratio and translates into the 

quantity of steam that is routed to the precooler. Because the highest pressure drop is 

assumed for conservative design of the vacuum system, a larger quantity of steam is 
assumed to be exhausted from the condenser at a pressure lower than that 

actually expected. This penalizes the 

vacuum system because it requires more 

work to be done in compressing the gases, 

Steam plus noncondensable gases resulting in predicted parasitics that are a 

few percent higher than expected. If the 

seawater overall pressure drop in the condenser is 

out lower than has been assumed for the nominal design, this penalty may be recovered. When the vacuum system is better 

defined, this issue will be considered in 

detail. 

5.5 Design Specifications 

Specifications for the heat exchangers are 

summarized in Table 5-3. Values of 

parameters derived from the specif ications are added for reference. 

5.6 Interfaces 

The interfaces between the heat exchangers and the rest of the NPPE are 

shown in Figure 5-1. All penetrations 

through the vacuum vessel walls and sealing of these penetrations are the responsibility of the structures task. Plumbing on 

either side of the vacuum vessel will be 

designed in the respective tasks. The seawater systems task will provide all plumbing for the main heat exchangers up to the 

flanges on the vacuum vessel walls. 

Plumbing inside the structure, to distribute and collect the warm and cold 

seawater and to remove the uncondensed 

steam and noncondensable gases, will be 

designed under the heat exchangers task. 

Steam interfaces inside the vacuum vessel 

are at the turbine inlet plenum and at the 

steam diffuser outlet. Seawater and gas 

Figure 5-6. plumbing to and from the desalinatedElemental water condenser will be designed in the 

section of heat exchangers task. Control signals will 

test unit be received from the control room and instrumentation signals will be sent to the 

control room. Electrical service is necessary inside the vessel mainly for lighting. Noncondensable gases Compressed air or electrical power will be 

needed inside the vessel to regulate the 

BAG0603005 water split between condenser stages. 

5.7 Schedule 

The schedule for the heat exchanger task at the time of writing, shown in Table 5-4, is 

compatible with the summary schedule shown in Figure 1-3. Constraints on the schedule 

for the heat exchangers stem primarily from the schedule for the vacuum vessel 

structure, which must be built before heat exchanger components can be installed. 

Table 5-2. Performance Predictions at 

Nominal Conditions (Summary) 

Evaporator 

Thermal effectiveness 0.9 (+0.03) 

Mass of steam produced 3.5 kg/s 

Noncondensable fraction 0.3% 

Seawater-temperature drop 3.5OC 

Hydraulic losses 

- Static 

No delays are envisioned because of heat exchanger 

components, because all 

supplies and equipment are 

currently available. The 

same is true for the water distributors, structured 

packing sections, and mist 

eliminator. 

5.8 Cost Estimates 

- Dynamic 

Steam temperature to turbine Og5 m, The estimated cost of spe- 21.8OC 

<5 Pa cific equipment for the heat Mist eliminator pressure drop exchanaer task is summarDirect-Contact Condenser 

Thermal effectiveness 

Vent ratio 

Percent steam uncondensed 

Seawater-temperature rise 

Pressure drop 

Hydraulic losses 

- Static 

- Dynamic 

0.86 (+0.03) 

0.8 

0.8% (-0.3%) 

5.Z°C 

120 Pa (-37 Pa) 

ized inuTable 5-5. Costs include purchase of material 

and labor to build and assemble the components, including piping from the walls 

of the dome. Costs do not 

include labor to install the 

heat exchanger components, 

because it is assumed that a 

team will be on site to install the structure. 

Developments made during the past decade in direct-contact heat exchangers have 

reduced not only their performance uncertainty but their cost, especially in the dome 

configuration. Notice that the cost of the heat exchanger equipment amounts to only 3% 

of the total equipment cost estimate for the HMTEAINPPE. 
Table 5-3. Heat Exchangers: Specifications and Related Parameters 

Evaporator 

Type 

Seawater flow 

Inlet-seawater temperature 

Steam generation rate 

Steam temperature 

Liquid loading 

Thermal effectiveness 

Active spout height 

Number of spouts 

Diameter of spouts 

Spacing between spouts 

Chamber width 

Related Parameters 

Evaporator planform area 

Inside diameter 

Outside diameter 

Steam generation velocity 

Spout-seawater velocity 

Outlet-seawater temperature 

Supply pipe diameter 

Water distribution pipe diameter 

Discharge pipe diameter 

Mist Eliminator 

Type 

Material 

Passage spacing 

Ratio of flow to planform area 

Planform area 

Distance above top of spouts 

Loss coefficient 

Steam velocity through mist eliminator 

Mist removal efficiency 

Direct-Contact Condenser 

Type 

Number of stages 

Packing 

Seawater flow 

Seawater remperature 

Steam consumption rate 

Steam inlet temperature 

Inlet noncondensable mass fraction 

Overall thermal effectiveness 

Steam loading to first stage 

Liquid loading of first stage 

Area ratio of stages 

Packing height of first stage 

Vertical spout 

0.5 m 

(12 + I) 

0.24 m (10 in.) 

21 m 

11 m 

17.7 m2 

4.27 m (14 ft) 

6.40 m (21 ft) 

10 m/s 

1 m/s 

22.5OC 

0.76 m (30 in.) 

0.61 m (24 in.) 

0.91 m (36 in.) 

Munters chevron 

Plastic 

19 mm (314 in.) 

78% 

28 m 

12 m 

10 

6.2 m/s 

90% removal for droplets of diameter 

greater than 40 um 

Column, structured packing 

2 

Munters CF-25060 

42 1 kg/s 

6.1 OC 

3.5 kg/s 

11.8OC 

0.3% 

0.86 

0.4 kg/rn2s 

38 kg/m s 

3 

0.61 m (2 ft) 
Table 5-3. Heat Exchangers: Specifications and Related Parameters (Concluded) 

Direct-Contact Condenser (Concluded) 

Packing height of second stage 0.91 m (3 ft) 

Interstage steam passage height 0.47 m (1.5 ft) 

Related Parameters 

Area of first stage 

Area of second stage 

Liquid loading of second stage 

Overall pressure drop 

Steam temperature at outlet 

Flow rate to first stage 

Flow rate to second stage 

Maximum height for water 

distribution tray 

Supply pipe diameter 

Discharge pipe diameter 

Outside diameter of turbine 

diffuser exit flange 

Inside diameter of turbine 

diffuser exit flange 

0.30 m (1 ft) 

0.61 m (24 in.) 

0.76 m (30 in.) 

4.11 m (13.5 ft) 

Surface Condenser for Desalinated Water Production 

Overall Configuration 

Condenser length 3.05 m (10 ft) 

Width 0.57 m (1.87 ft) 

Effective height (per section) 0.43 m (1.4 ft)2 

Heat-transfer area (per section) 31.1 m (334 ft 4 

Steam-side flow area 0.95 m 90.3 ft 1 

Water-side flow area 0.046 m (0.5 ft ) 

Number of water channels 12 

Number of steam channels 13 

Steam Channels 

Channel spacing 25.4 mm (1.0 in.) 

Fin thickness 0.8 mm (0.032 in.) 

Fin density section 1 29.5 fpm (0.75 fpi) 

2 39.4 fpm (1.0 fpi) 

3 59.2 fpm (1.5 fpi) 

Water Channels 

Extrusion width 

Number of flow channels 

Channel size a 

b 

Radius of curvature 

Wall thickness 

Equivalent diameter 

462.3 mm (18.2 in.) 

28 

12.7 mm (0.5 in.) 

12.7 mm (0.5 in.) 

4 mm (0.16 in.) 

2.54 mm (0.1 in.) 

13.3 mm (0.523 in.) 

Correct heights and nominal diameters of pipes are given in ft/in. 
Table 5-4. Projected Schedule for Heat Exchangers 

Complete concept of direct-contact condenser 

water distributor and flow-split control 

Complete design of surface condenser 

Complete preliminary design of heat exchanger 

components 

Complete final design of heat exchanger 

components 

Fabricate water distributors, packing sections, 

and internal partitions 

Install supports in dome 

Complete installation inside the dome 

(depends on availability of structure) 

Install instrumentation 

Install turbine diffuser and mist eliminator 

Install surface condenser 

Initiate HMTEA tests 

February 1990 

March 1990 

June 1990 

September 1990 

March 1991 

April 1991 

May 1991 

June 1991 

June 1991 

July 1991 

July 1991 
Table 5-5. Breakdown of Heat Exchanger Costs 

Subtotal 

Evaporator/Mist Eliminator 

Water distributor 

Mist eliminator 

Supports 

Direct-Contact Condenser 

Water distributor and controls 

Packing, fabricated 

Internal supports and partitions 

Shipping 

Surface Condenser 

Drawings/Installation 

Drawings 

Installation supplies 

Test Instrumentation 

$20K 

$1 OK $30K 

$40K 

Total $300K 
CHAPTER 6.0 

ANCILLARY SYSTEMS 

6.1 Objective 

Responsibilities of the ancillary systems task a r e t h e vacuum compression system (VCS), 

the electrical system, the instrumentation and control system (ICS), and other 

miscellaneous support systems. The objective of this task is to develop these systems for 

the HMTEAINPPE at the STF. 

6.2 Goal 

The task goal is to develop ancillary systems for the NPPE that consume no more than 

44 kWe of power at the NPPE design point. 

The key ancillary component affecting this goal is the VCS. A target of 20-kWe power 

consumption for the VCS was established early in the program based on performance predictions of an untested, large-scale plant. Later assessments of the overall NPPE power 

budget indicated that a VCS power consumption target of 40 kWe would still allow the 

experiment's overall net power goal of 40 kWe to be met. The ancillary systems task 

team has invested the majority of its conceptual design effort toward developing of a 

high-efficiency VCS to meet this 40-kWe target. 

Other components of the ancillary system do not have unusual requirements. Standard 

engineering practices are being used in designing the electrical system, the ICS, and the 

miscellaneous support systems. A total power consumption target of 4 kWe was set for 

these last three areas. 

The members of the task team at the time of writing are 8. Parsons of SERI, R. Davis of 

PICHTR, and T. Rabas of ANL. 

6.3 Vacuum Compression System 

The VCS must continuously remove noncondensable gases and residual steam from the 

condenser to maintain system operation. Noncondensable gases are present in the incoming warm and cold seawater. At subatmospheric heat exchanger pressures, the dissolved 

gases are supersaturated, and between 80% and 100% of the noncondensable gases come 

out of solution. In addition, a small amount of air will leak into the vacuum enclosure 

from the atmosphere. 

Two methods of gas compression were proposed for OC-OTEC systems. The first is a 

multiple-stage mechanical compression train. The second uses discharge seawater to 

hydraulically compress after an initial mechanical compression stage. The hydraulic 

compression option has potential for reducing the overall VCS power consumption but 

was eliminated from consideration for the HMTEAINPPE design because of the research 

and development effort required and the preference for adapting existing and proven 

hardware into the VCS design. 

In the multiple-stage mechanical compression train, interstage gas coolers substantially 

reduce the power consumption. Cooling the gas before entry into the next compression 

stage minimizes the work required to compress a given mass of gas and is necessary so 

the high-temperature limits of the compressor components are not exceeded. The 

intercoolers can also be used to condense a large portion of the residual steam, therefore 
reducing the gas mass flow rate that 

must be processed by the downstream 

compression stages. The VCS also includes compressor motors and a selfcontained lubrication system. The general staged layout is illustrated in Figure 

6-1. 

6.3.1 Design Approach 

A simple model was developed to calculate power consumption of a staged compression system using efficient intercoolers. Initial efforts included 

exercising the model to define key performance parameters. The results are 

summarized in Table 6-1. 

A major effort by SERI and PICHTR personnel was mounted to define commercial compressor efficiencies, sources 

of inefficiency, and possible improvements. 

Using existing equipment was found to be 

impractical, but the potential for improvement using existing technology and 

integrating low-loss components was 

identified (see Section 6.3.3 for details). 

The conceptual design effort culminated 

in the development of a statement of 

NClsteam 

X from condenser 

B 

S Cooling 

water in 

I Precooler 

I I I I Cooling 

-water out 

I 

I 

I 

Power in 

I 

I 

I 

I 

Controlldata 

signals I 

I 

I 

I 

I 

I 

Compressor 

I 

I I 

------- I work (SOW) for preliminary design of a 

VCS. The SOW document defines oper- Remaining NCIstearn 

exhaust to atmosphere atinn conditions for a VCS usinn a directcon&ct precooler. The SOW Gas issued Figure 6-1. Conceptual design of 

by PICHTR in September to five com- vacuum compressor system 

pressor designers and manufacturers, 

and the preliminary design contract was forwarded to Barber-Nichols Engineering in 

early November 1989. Details of the SOW and contract price are in Section 6.3.8. 

6.3.2 Inlet Conditions 

The overall plant system operating parameters and the condenser performance determine 

the VCS inlet conditions. Table 6-2 lists the basis for arriving at the requirements for 

the VCS design. 

The final parameter needed to size the vacuum system is the amount of steam leaving 

the direct-contact condenser. Instrumentation accuracies of HMTSTA tests result in 

significant uncertainties in condenser outlet steam flow rate. To ensure a more limited 

range of inlet flow rates to be met by the VCS, a precooler was included in the design. If 
Table 6-1. Key Performance Parameters for a VCS 

Parameter Action 

Compressor efficiency Engineering development effort required. 

Inlet conditions None. Inlet conditions are determined by overall system 

considerations and heat exchanger performance. 

Precooier performance None. Performance predictions are based on HMTSTA 

direct-contact and surface condenser test results. 

Intercooler performance Experimental verification of direct-contact and surface 

condenser predictions at high noncondensable gas concentrations is recommended; these tests are 

recommended to be carried out at the SERI facilities. 

Table 6-2. Basis for Defining the Requirements for the VCS Design 

Parameter Value Note 

Warm-seawater 620 kg/s ?lo% for NPPE tests, +lo%, -50% for heat 

flow rate exchanger HMTEA tests. 

Cold-seawater 

flow rate 

Cold-seawater plant 

delivery temperature 

420 kg/s Same as above 

Condenser-seawater 11.2OC --- 

outlet temperature 

Percentage dissolved 

noncondensable gas 

release 

100% In heat exchangers with no predeaeration, 

VCS should accommodate variabilities down 

to 80%. 

Atmospheric leakage 1 g/s --- 

Diffuser outlet steam 1 1.8"C ?0.5"C design variation. 

temperature 

Normal condenser pressure 120 Pa 100-150 Pa 

drop 
it is determined during heat exchanger checkout tests that the condenser is performing 

at the expected level, the precooler may be removed from operation. To maximize the 

efficiency of the precooler, a countercurrent direct-contact device using cold seawater 

(at 6.1 "C) was selected. 

The VCS inlet operating requirements are listed in Table 6-3. This table covers a range 

of inlet pressures and a range of noncondensable gas flow rates to encompass the 

potential range of condenser pressure loss and gas liberation. These conditions are downstream of the precooler and assume a precooler outlet approach temperature (the difference between inlet water temperature and outlet gas temperature) of 0.5OC. This 

approach temperature is slightly greater than analytic code predictions. Reducing the 

assumed approach temperature to zero does not greatly affect the steam flow rate., 

Table 6-3. Inlet NPPE Operating Requirements In addition to the operating 

for the Preliminary VCS Design conditions for the steady-state, 

Contract full-system test in which net 

power will be determined, ' the 

Inlet Steam Flow Rate (g/s) specification for the prelimPressure as a Function of Inlet Pressure inary design contract included 

(Pa) pump-down conditions and heat and Noncondensable Gas Flow Rate exchanger checkout test condiNoncondensable Gas Flow Rate (g/s) 

tions. 

6.3.3 Compressors 

This section discusses the results of investigating existing 

commercial compressors for 

use in the HMTEAINPPE and 

possible performance improvements of compressors using 

existing technology. 

Centrifugal and positive-displacement compressors were the two existing commercial 

types investigated. The third type, the axial compressor, was eliminated from consideration for the following reasons: 

Existing axial commercial hardware is designed for higher volumetric flow rates. 

Extensive development would be required for smaller sizes. 

Lower pressure rise per rotor results in numerous stages and high costs. 

We contacted 13 manufacturers of positive-displacement compressors. These machines 

are typically used at vacuum levels compatible with OC-OTEC, but their typical 

applications are not overly concerned with efficiency. The primary design advantages of 

existing positive-displacement compressor hardware are reliability and flexibility of 

operation. Existing positive-displacement integrated multiple-stage systems typically 

run with only two or three stages. Resulting stage pressure ratios are higher than 

desirable for our application, but lower pressure ratio operation is possible. 

Manufacturers suggested rotary-lobe and rotary-vane machines, with liquid ring pumps 

acting as the final stage. Rotary-screw types were also mentioned, but the high pressure 

ratios of these machines made them less attractive. Compressor stage efficiency 

numbers were difficult to obtain from manufacturers, but inferred efficiencies were 

found to range from 25%-36%. These figures were partially confirmed by measurements 

on an existing rotary-lobed blower in the SERI heat- and mass-transfer laboratory, which 

had an efficiency of 32% over a wide range of operating pressures. 
Two system performance predictions were done with the inferred compressor stage efficiencies to evaluate the feasibility of using existing positive-displacement hardware. 

Operating conditions for these runs were taken from a nominal point in the preliminary 

VCS design specification. The first prediction corresponds to using existing rotary 

positive-displacement compressor stages at typical pressure ratios. The first two stages 

are rotary-lobed blowers with an overall stage efficiency of 3296, and the final stage is a 

rotary-vane blower with an efficiency of 28%. The power calculation assumes that efficient intercoolers are used to reduce the steam mass flow rate. Use of intercoolers is an 

addition to existing positive-displacement system designs, but it would be possible to 

purchase individual compressors and integrate the stages with intercoolers. Table 6-4 

summarizes results. The compressor power consumption projection of 62 kW clearly 

indicates that use of existing positive-displacement compressor hardware a? typical 

pressure ratios is not a viable option. 

The second prediction regarding existing positive-displacement compressor performance 

assumes that the pressure ratios can be reduced and that the highest inferred efficiencies 

are applicable at these lower pressure ratios. Again, efficient intercoolers are used. 

Results are shown in Table 6-5 for a six-stage equal-pressure-ratio compression train. 

The above power estimates are for the compressors only. Pumping lubrication oil or 

cooling water would require additional power. Based on these calculations, existing 

positive-displacement pumps are unacceptable for use in the VCS. 

We contacted seven manufacturers of centrifugal compressors. Existing centrifugal 

compressors are not designed to operate at low pressure levels required. Machines are 

available that could accommodate the gas volumetric flow; however, because they are 

designed for higher density gases, the drive systems and bearings are sized for large 

power requirements and lower operating speeds. The mechanical losses associated with 

existing centrifugal compressor designs operating at higher speeds are excessive for our 

low-density gas application. As an illustration, Atlas-Copco has an existing centrifugal 

wheel with approximately 77% gas compression efficiency. Acting as a first stage, gas 

compression would require around 8 kWe. However, the oversized mechanical system 

results in losses that increase motor power to 34 kWe and decrease compressor efficiency 

(not counting the motor) to less than 20%. Therefore, existing, off-the-shelf centrifugal 

compressors are also unacceptable for use in the VCS. 

Table 6-4. Compression System Power Projection Using Existing PositiveDisplacement Compressor Hardware at Standard Pressure Ratios 

(Total compressor power required = 6 1.8 kWe) 

Volumetric 

Inlet Outlet Steam Flow Rate 

Compressor Pressure Pressure Pressure Flow Power 

Number Ratio (kPa) (kPa) (g/s) (kwe) (m3/s) (cfs) 
Table 6-5. Compression System Power Projection Using Existing PositiveDisplacement Compressor Hardware at Lower-Than-Standard 

Pressure Ratios (Total compressor power required = 44.4 kWe) 

Volumetric 

Inlet Outlet Steam Flow Rate 

Compressor Pressure Pressure Pressure Flow Power 

Number Ratio (kPa) (kPa) (g/d (kwe) m3 (cfs) 

1 2.1 1.2 2.5 51.54 17.82 6.865 14536 

2 2.1 2.5 5.2 22.99 9.75 1.823 3860 

3 2.1 5.1 10.9 5.60 5.06 0.436 924 

4 2.1 10.8 22.8 2.16 4.18 0.167 353 

5 2.1 22.8 48.0 0.94 3.87 0.072 153 

6 2.1 48.0 101.3 0.43 3.75 0.033 70 

44.4 

Conclusions drawn from centrifugal and positive-displacement compressor manufacturers 

include the following: 

a Existing compression stage hardware has excessive mechanical losses that reduce 

overall efficiency to an unacceptable level for our application. 

a Gas compression efficiency appears to be adequate, indicating that existing rotor 

designs can be used. 

a Improvements in mechanical losses appear possible by down-sizing centrifugal drive 

systems and integrating available low-loss bearings and seals with either type of 

compressor. 

a Engineering design efforts are required to develop a machine with low mechanical 

losses and to design appropriate intercooler staging integration schemes. 

6.3.4 Gas Condensers and Coolers 

As discussed previously, intercoolers between compression stages are key components in 

the VCS. The precooler between the direct-contact condenser and the vacuum compression system will also be discussed in this section. 

Commercial multistage compression trains use intercoolers to reduce the temperature of 

the gas. For these gas coolers, cooling water flow rates are relatively low, and heat 

exchange is easily accomplished as a result of large available temperature differences. 

A conventional shell-and-tube heat exchanger with the gas on the shell side and water on 

the tube side is typical. 

For our application, both precooler and intercooler act as steam condensers. The 

primary design consideration is to reduce the steam mass flow by condensation in the 

presence of high concentrations of noncondensable gases. The concentration of noncondensable gases is about 25% to 38% (by mass) at the compressor train inlet, increasing to 

greater than 95% at the outlet. This concentration is much greater than typical power 

plant condensers and OC-OTEC condensers. 
Table 6-6. Summary of Gas Condenser/Cooler Types Considered for the Precooler 

and Intercooler Designs of the HMTEAjNPPE VCS 

-- -- 

Cooler Type Advantages Disadvantages Action 

Direct contact Lowest approach Siting complications Use direct contact 

temperature Flooding problem for precooler, 

Lowest cost Complex piping carry as an 

Minimal corrosion arrangement option for 

consideration Potentially higher intercoolers 

Low gas-side cooling-water 

pressure drop parasitics 

Plate-Fin Low gas-side Potential higher cost Carry as an option 

pressure drop than direct contact for intercoolers 

No siting problems Higher approach 

Simple piping temperature than 

arrangement direct contact 

Conventional Proven cooler type Higher percentage Carry as an option 

shell and tube Typical choice for noncondensable for intercoolers 

standard VCS gas concentration 

raises performance 

doubts 

Highest approach 

temperature 

The three gas condenser/cooler types considered for the precooler and intercooler 

designs were 

A conventional shell-and-tube heat exchanger as described 

A direct-contact condenser/cooler type 

A plate-fin condenser/cooler type. 

Table 6-6 summarizes the advantages and disadvantages of the three options and the 

status of the precooler and intercooler designs. 

Another gas cooler design issue is the type of cold water used. Options include 

Coldest seawater from the supply system at 6.1°C 

Effluent cold seawater from the condenser discharge at 11.3OC 

Closed-loop, treated tap water in a liquidlliquid heat exchanger cooled by either of the 

first two types of cold seawater. 

The coldest seawater is most useful for the lowest pressure gas coolers, where the 

concentration of water vapor is highest. The use of the coldest seawater must be minimized because the total plant flow rate is limited. In addition, corrosion associated with 

seawater must be considered in the design. 
Main condenser effluent water is cold enough for many cooling needs. Corrosion problems are minimized by using a closed-loop system, and the gain of several degrees 

Celsius in a liquidlliquid heat exchanger may not be critical. 

The precooler is immediately adjacent to the first-stage compressor outside the dome. 

This will limit the effects of ambient heating of the low-density gas in the large diameter pipe. Piping from the main condenser through the vacuum dome to the precondenser 

cooler is 0.6 m in diameter. Assuming 20 m of pipe with six 90-deg bends (near 100-m 

total equivalent length) results in a piping pressure loss of around 15 Pa for the high flow 

rate conditions of Table 6-3. 

The precooler performance is important because it determines volumetric and mass flow 

rate for the first-stage and hence the largest compressor. A direct-contact device with 

6.1°C seawater was selected to minimize steam flow into the compressor. An approach 

temperature of 0.5"C was used to develop the VCS design specification. Using a closedloop water supply or a surface condenser results in a larger steam flow rate to the first 

stage. Using effluent water is not reasonable because even worst-case main condenser 

performance results in steam temperatures much lower than 11.3OC. 

SERI's direct-contact condenser code was used to check the precooler performance. A 

gas pressure drop of less than 50 Pa was predicted. Approach temperatures less than 

0.5"C were projected to be achievable, but up to 25-kg/s cold-water flow may be 

required for worst-case main condenser exit conditions. Up to 1.5 kWe of additional cooling-water pumping power may be required if the precooler is not located at barometric 

height. Further design work optimizing precooler parameters may result in lower worstcase water flow rates. 

Further investigation is required before decisions are made regarding the type of cold 

water and the type of cooler used for each intercooler. Successively decreasing steam 

flow rate allows higher liquid loadings to accommodate higher noncondensable gas concentration without increasing water flow rates and pumping power. For higher pressure 

downstream intercoolers, the gas pressure drop and gas exit approach temperature do not 

have a great effect on VCS power, so a surface-type heat exchanger may be more appropriate. Using effluent cold water instead of 6.1 "C seawater for the intercoolers results 

in a predicted increase of only 1-2 kWe in VCS power, with more than 50% of the 

increase resulting from the performance of the first intercooler. Intercooler water flow 

rate is expected to be between 10 and 25 kg/s, but it is not as critical as precooler water 

flow if main condenser effluent seawater is used. 

The detailed design of the precooler and intercoolers will be coordinated by the VCS preliminary design contractor with assistance as required from SERI and ANL. Precooler 

and intercooler costs are estimated to be $60,000 over and above the cost of the 

preliminary and final design contract. 

6.3.5 Compressor Drivers 

Three-phase AC motors were selected to drive the compressors. The power supply corresponds to the turbine/generator output, so no transformers are required. A variablefrequency drive may be used if compressor rotational speed is selected as a control 

mechanism (a typical variable-speed-drive efficiency is 95%). 

Some past commercial-scale open-cycle plant designs have used compressors directly 

driven by the turbine. Direct-drive systems avoid power losses from motor and generator 

inefficiencies, but the compressors must be located near the turbine shaft and compressor control may be more complicated. 
The motors will be selected by the compressor manufacturers. This will allow starting 

and steady-state characteristics to be selected to match compressor requirements. 

An air-cooled motor may have a design point efficiency of near 90%. Typical data from 1 manufacturers show a few points' increase in efficiency for water-cooled motors; 

therefore, the compressor manufacturer may consider them. The increase in coolingwater pumping power and the added cost of cooling-water circulation equipment must be 

considered in the evaluation. 

6.3.6 Other Vacuum Compressor Support Equipment 

i The VCS will require lubrication for moving parts. The detailed design of a lubrication 

system is included in the VCS preliminary design contract. The contractor will select a 

lubrication system appropriate for their VCS design. 

Lubrication system requirements may be very different for positive displacement and 

centrifugal compressor types because of rotational speed and sealing requirements. A 

closed-loop oil-circulating system including a pump, oil reservoir, oil filter, and other 

elements may be required. Another option is sealed, packed bearings, which will be 

replaced at regular maintenance intervals. Water contamination of the oil must be considered if oil comes in contact with the process gas. 

Oil-sealed positive-displacement pumps will use more oil than any other option and will 

require more oil handling and processing equipment. The lubrication system design is 

standard practice, and an oil pumping power requirement of less than one kWe is 

expected. 

Control of the VCS over the range of expected steady-state operating conditions and during vacuum chamber pump-down is also an issue. Mechanisms for control may be different for centrifugal and positive-displacement machines. For a given rotational speed, 

centrifu a1 machines develop a nearly constant head (or pressure ratio if flow rate is 

constant 7 as inlet volumetric flow is varied. Positive-displacement machines, on the 

other hand, have a nearly constant capacity over a wide range of head (see Figure 6-2). 

Centrifugal machines also have a defined lower limit of inlet flow called the surge 

limit. If the machine operates near the surge limit, unstable flow, caused by back-flow 

into the rotor, induces vibrations that can destroy the compressor. Control of either 

type of machine is feasible through a variety of schemes, including 

a Rotor speed control by a variable-speed drive 

a Gas bleed-in or recirculation for low-flow conditions 

a Parallel flow paths (full or partial compressor trains handling a fraction of the total 

flow at highest flow rates). 

, Contractors will define the specific control schemes in the preliminary design. 

6.3.7 Conceptual Design Performance Prediction 

Two predictions of VCS power consumption are important. The first uses expected condenser and system performance levels; the second uses worst-case inlet conditions from 

the vacuum specification of Table 6-3. 

For both cases, an overall compressor stage efficiency of 60% is used (90% motor, 94% 

variable-speed drive, and 72% compressor). Effluent water (11.3OC) is used as the 

cooling sink for a closed fresh-water loop, resulting in an intercooler water supply 
temperature of 13.5OC. For these predictions, four compression stages are 

used with equal pressure ratios in each. 

Other assumptions are that 100% of the 

noncondensable gases are released and 

the pressure loss between the compressor 

stages is 200 Pa. Power consumption 

may be slightly reduced by increasing the 

number of stages and reducing the pressure ratios of the first two stages. 

The expected perf ormance prediction 

uses a turbine exit temperature of 

11.8"C and a total pressure drop from 

the diffuser to the first compressor inlet 

of 150 Pa. An approach temperature difference of -O.l°C was used (approach 

can be negative because of the boiling 

point elevation of seawater), resulting in , 

an inlet steam flow rate of 37.3 gfs. 

Power consumption for the compressors 

was calculated to be 26.5 kWe. 

The worst case from the preliminary design specification uses a higher inert gas 

flow due to higher seawater flow, a 

lower compressor inlet pressure, and an 

approach temperature of 0.5OC. The resulting inlet steam flow rate is 52.4 

g/s- Power consumption for the 

compressors was calculated to be 3 1.8 

kWe. 

Positive displacement \ 

0; 

I I 

50 100 

Capacity (%) 

Figure 6-2. Pressure-volume diagram 

for centrifugal and positive-displacement 

types of machines. Capacity is usually 

specified in terms of inlet CFM (at 

inlet conditions) and head in feet 

or pressure in PSI for compressors 

and blowers. For fans, however, 

the pressure is usually specified 

in inches of water gauge. 

Cooling-water circulation for the gas precooler and intercoolers is expected to consume 

a total of nearly 5.5 kW if barometric-height direct-contact devices (or equivalent 

performance/head loss surface condensers) are used. 

Vacuum pump lubrication requirements will be fully defined by the preliminary design 

contractors, but power to circulate oil is expected to be under 1 kWe. The lubricating oil 

may need to be cooled in a liquidfliquid heat exchanger. Oil cooling-water pumping 

power is estimated to be 0.5 kWe. 

Cooling water may be required for the compressor motors, but any water pumping power 

requirement should be more than offset by reductions in compressor power due to motor 

efficiency increases. 

Total VCS power consumption predictions are 34 kWe if expected performance is 

achieved and 39 kWe if worst-case inlet conditions are used. 

A preliminary layout for the VCS indicates a 6-by-6-m (20-by-20-ft) footprint. 

Compressor stage dimensions were estimated using existing appropriately sized 

volumetric-flow, positive-displacement rotary blowers. 
6.3.8 Preliminary Design Contract 

The conceptual design of the VCS defined inlet conditions, identified limitations of existing equipment, and made power consumption projections based on improved compressor 

performance. The preliminary design contract involves a compressor manufacturer confirming these improved compressor performance projections and completing the VCS preliminary design. The preliminary design contract is Phase I of a two-phase effort. 

Phase I1 includes final design, construction, and testing of the Phase I design. Total cost 

of both phases is estimated by Barber-Nichols Engineering to be $225,000. 

6.3.9 Remaining Issues and Uncertainties 

I 

Remaining issues and uncertainties that must be addressed in the VCS preliminary design 

are to 

Confirm that compressor mechanical losses will be low and that an overall power 

consumption of around 40 kWe can be achieved. 

Experimentally verify direct-contact and surface condenser code predictions of precooler and intercooler performance at high noncondensable gas concentrations. 

a Work with the VCS contractor in the final selection and design of the precooler and 

intercooler. 

Resolve pipe routing and component placement issues to minimize pressure loss, maximize performance, and prevent compressor flooding. 

The schedule of key events and deadlines for the VCS is included in the summary section 

of the report. The primary constraints are the completion of the compression system 

Phase I and Phase I1 contracts before beginning heat exchanger checkout tests. 

6.4 Electrical System 

The complete electrical system for the NPPE includes the generator and associated 

switchgear, wiring and bus connections, interface with the existing NELH grid system, a 

resistive load bank to dissipate generated power, motor control centers, emergency 

power supplies, and power monitoring devices. 

The requirements for the generator, associated switchgear, and load bank are primarily 

dictated by turbine start-up and control characteristics; therefore, the specific design of 

these components is included as part of the turbine design. Start-up scenarios and controls, however, do impose conditions on the remaining electrical components, so close 

' coordination is necessary. 

6.4.1 Design Approach 

The electrical system design is similar to existing power plant designs. Standard engineering practice and hardware are expected to be used for all aspects. PICHTR has primarily undertaken the ancillary electrical system design. As part of that effort, PICHTR 

contracted with a local (Honolulu) electrical engineer, H. Miura, for much of the detailed 

work. General electrical requirements were gathered from the systems task team and 

other component design teams. From these requirements, PICHTR and Mr. Miura developed one-line diagrams, equipment lists, and cost estimates. 

6.4.2 System Requirements and Options 

The ancillary electrical system must supply power for all plant needs and provide protection from faults during start-up, operation, and shutdown. During normal operation, the 
electrical system must be configured to allow determination of net power production 

from the NPPE. 

The general operational scenario is that the turbinelgenerator will use the existing 

NELHJHawaiian Electric Co. (HELCO) power grid for excitation and a synchronization 

standard. All HMTEAINPPE power needs will be met by HELCO during start-up. The 

generator may be used as a motor to bring the turbine up to speed. Once the turbine is 

up to speed and synchronized, the plant loads essential to operation will be switched over 

to OTEC power. 

If a power outage occurs in the NELHIHELCO system, the experiment will shut down 

because excitation is provided by the grid. During any power loss or similar shutdown, 

emergency power must be provided to operate critical components. Two types of emergency power needs were identified. The first cannot tolerate any loss of power and 

therefore must be backed up by a battery-charged uninterruptable power supply (UPS). 

Essential data-acquisition and control hardware will be on the UPS. The second type of 

emergency power need can tolerate a short delay in power. To reduce the size and coss 

of the UPS, this type of need is met by an emergency engine-driven generator. NELH 

personnel have indicated that excess capacity from NELH emergency generators could 

provide the needed HMTEA/NPPE backup with the provision of additional switchgear. 

Critical power users that may be able to tolerate a momentary delay in power include 

Main and backup lubrication oil and cooling-water pumps (turbine and vacuum system) 

Turbine scavenge oil pump (if shutdown does not include breaking the vacuum in the 

seal area) 

Air compressor for controls (if required). 

In addition, some uses such as emergency lighting and other safety-dictated needs should 

be on the emergency power system. Preliminary design will include a check of tolerable 

power delays and emergency power response time to ensure compatibility. 

Electrical power needs were divided into two categories, each handled by its own switchboard. Plant loads provide for essential operations and are included in the net power 

determination. Building and support loads are not a part of the power cycle and are not 

included in the net power determination. Plant cycle loads include 

Seawater supply and discharge pumps 

Vacuum compressors 

Cooling-water circulation pumps 

Lubrication oil circulation and scavenge pumps 

Backup pumps 

Instrumentation and controls (including air compressor if needed for controls). 

Building and support loads include 

Support building receptacles and lighting 

e Air-conditioning and ventilation (including any provisions for use of seawater cooling) 

Desalinated-water pump 

Tool air compressor 

Crane or hoists. 

Two options were evaluated during this conceptual design phase for net power distribution. The first option routes net power to a power-dissipating load bank, and the second 

option adds the ability to send power to the utility grid. Cost estimates indicate an additional $7,700 would be needed for the option to send power into the grid. 
Discussions held with HELCO and NELH personnel resulted in strong support for the netto-grid option. The systems integration task team concurred, so remaining net power 

will be routed to the main NELH system to offset a portion of HELCO-supplied power for 

other NELH users. No OTEC power will actually reach the main HELCO grid because 

NELH has a continuous need for 80-95 kWe. 

6.4.3 Cost Estimates 

The following summarizes the ancillary electric system cost estimate (as included in the 

ancillary systems task), prepared by H. Miura. It includes material and labor. An 

estimated 20% contingency is not included in the summary costs (see Table 6-7). 

Electrical equipment associated with the turbine, including generator switchgear, a control panel, the load bank, switching, and relays, was estimated to cost $207,000. It is 

included in the turbine portion of the cost estimate. 

6.4.4 Remaining Issues and Uncertainties 

No critical issues or uncertainties were identified in the ancillary electrical system conceptual design. Refinement of the design is needed as requirements are better defined 

and revised by other task areas. Preliminary design will require close coordination with 

the turbine contractor to ensure compatibility with turbinelgenerator controls and 

switchgear. 

6.5 Instrumentation and Control Systems 

Only a small effort was mounted in designing instrumentation and control systems. Plant 

control, dynamics, instrumentation, and data-acquisition design are expected to use 

standard engineering practices and available equipment. Much of the information 

required to define a comprehensive plant control system will become available after partial completion of the preliminary design. Two key analyses are planned as part of the 

control definition. The first is a plant dynamic modeling effort to define required control response times and possible control mechanisms. The second is a failure modes and 

effects analysis to ensure that controls address plant safety issues. These efforts will be 

coordinated with the systems integration task team and the seawater systems task team 

(which is responsible for safety). 

For conceptual design, general control and instrumentation philosophy was defined, firstcut instrumentation to monitor the plant and define component performance was 

identified, and an order-of-magnitude cost 

Table 6-7. Cost Estimate for Ancillary estimate based on typical power plant 

Electric System control and monitoring systems was 

obtained. Ancillary Estimated 

Electric System Cost 6.5.1 Design Approach 

Motor control centers $75,800 

Wiring 97,000 The plant controls and instrumentation 

Switchboards and relays 33,000 design will accommodate the following 

UPS 26,000 constraints: 

Transformers 5,100 

Miscellaneous 54,100 Numerous startlstop cycles expected 

during initial checkout, with operation 

Total $29 1,000 for 8-12-h periods 
a Continuous operation (24 hlday), with one operator on duty for an extended period of 

up to six months with minimal shutdown (short periods of unattended operation) 

a Semiautomatic control, with manual backup and overrides and automatic failure detection and shutdown 

a Individual components designed to be checked out and run without the full system 

operating 

An experimental facility that will be instrumented to allow component and system performance to be monitored, recorded, and analyzed. 

The turbine, vacuum system, and seawater pumps will have their own local control panel 

with an interconnect to a main control room. Local and main control and data-acquisition systems (DAS) will be modular to allow expansion and adaptation as experimental 

needs develop. All key valves and motor controls will be fitted for remote actuation. 

Much of the detailed design of the control and instrumentation system will be performed 

by consultants (with power plant experience), equipment manufacturers such as 

Foxborough or Rosemount, or both. 

Instrumentation required to monitor component and plant performance will be specified 

by each component design team. For example, the seawater flowmeters will be specified 

by the seawater systems task team. For these instruments, a standard output signal of 

0-10 V or 4-20 mA is preferred. When candidate instruments are identified, the ancillary 

design task team will coordinate the evaluation of overall probable measurement errors. 

6.5.2 Configuration 

Detailed design activity has not begun. The instrumentation and control needs for the 

turbinelgenerator are better defined than those for other components because its 

preliminary design is complete. The MTI Phase I report provides more information. 

Table 6-8 shows the initial inputs from the task leaders on the channel requirements for 

the performance monitoring instrumentation. 

An overall DAS limit of 128 channels was prescribed by the systems integration task 

team. The distribution of instrumentation and control channels among the various task 

areas will clearly need to be looked at in the preliminary design. Cost estimates will be 

based on this 128-channel limit. 

Table 6-8. Instrumentation and Controls Needed 

Some data will be logged or for the HMTEA/NPPE 

displayed on dedicated component control systems and 

DAS only; other channels will Channel 

also be routed to the main Subsystem Requirements 

control panel. Turbine and structures 5 0 

Power meter options for monitoring electricity production 

and consumption were identif ied. Specified accuracies 

range from 0.1% to 0.2% for 

Heat exchangers 

Seawater systems 

Vacuum compression system 

Ancillary electrical system 

Miscellaneous ancillary systems 

solid-state devices. Total 214 
Other instrumentation and control needs include 

Protection and safety 

Generator excitation 

LoadJspeed control for turbines, compressors, and pumps. 

The complete configuration of hardware for control and data acquisition remains to be 

defined. 

PICHTR and H. Miura have prepared a draft control room layout including electrical 

switchboards, motor control centers, and the generator control panel. The layout indicated a plan of 4.9 m by 6.1 m. Ergonomic layout and electrical interference issues 

remain to be addressed. 

6.5.3 Cost Estimate 

This first-cut evaluation of DAS needs is based on a channel count of 128. A PC-based 

system may be used because many flexible configurations are available. The overall DAS 

(all hardware and software) cost estimate is $75,000. The cabling and instrumentation 

cost estimate is $175,000, or $1,40O/channel. This figure was arrived at after consulting 

with instrumentation manufacturers and examining typical power plant costs. The overall control and instrumentation cost estimate is $250,000. 

6.5.4 Remaining Issues 

Control and plant safety issues will be addressed in the upcoming plant dynamicsJcontro1 

study and failure modes and effects analysis. 

No remaining design problems were identified or are expected that would require further 

development of the data-acquisition and control memory and processing units. 

Performance monitoring instruments were identified by the turbine preliminary design 

contract. Other instrumentation is expected to be similar to HMTSTA instruments with 

two exceptions. Temperature measurements for the HMTSTA tests were performed with 

platinum RTDs. Drift and frequent calibration needs make these instruments unsuitable 

for longer term HMTEAJNPPE tests. Industrial thermistors were identified as a possible 

solution, but further investigation is necessary to prove suitability. 

HMTSTA pressure measurements were mostly obtained using capacitance diaphragm 

transducers. Strain-gauge transducers were originally installed but were not able to 

withstand the seawater environment. Posttest, it was recognized that the capacitance 

diaphragm transducer measurement errors were much larger than thought because of 

temperature- and time-dependent stability. Suitable differential and absolute pressure 

transducers and appropriate installation and operational guidelines to minimize 

temperature- and time-dependent errors need to be identified for the HMTEAINPPE. 

Error analysis for all HMTEAJNPPE instruments and derived performance parameters 

remains to be completed. 

Reaeration of the discharge seawater stream may be required for environmental reasons; 

such needs remain yet to be explored, and power penalties associated with it are yet to 

be assessed. 
6.6 Miscellaneous Ancillary Systems 

The miscellaneous ancillary systems include cooling-water circulation systems, ventilation and air-conditioning (in conjunction with the structures task effort), and desalinatedwater systems (downstream of the surface condenser). 

6.6.1 Design Approach 

The miscellaneous ancillary systems design lags behind the main process because 

requirements are based on support needs. When the needs are defined, the process is 

straightforward because we do not expect any unusual designs or major effects on the 

main NPPE goals. The only miscellaneous system that affects net power is cooling-water 

circulation. Because supplying water for lubrication oil, vacuum system gas condensers 

and coolers, and possible motor cooling is essential for operation, pumping power is 

accounted for in net power determination. Other miscellaneous systems do not affect 

the net power. The designs for all miscellaneous ancillary systems are based on standard - 

engineering practice. 

6.6.2 Cooling-Water Circulation 

Cooling-water circulation is divided into uses that affect net power and supporting uses 

that are not essential for plant operation. Essential plant water systems service the 

following needs: 

VCS intercoolers 

Vacuum system and turbine lubrication oil coolers 

Vacuum system motors (an option) 

0 Control system air compressor. 

Supporting water systems service the following needs: 

Tool air compressor 

o Seawater air-conditioning loop (an option). 

Cold seawater directly from the supply system (at 6.1°C) or condenser effluent seawater 

(at 11.3"C) will be used as the heat sinks. Use of 6.1°C seawater is to be minimized 

because of the power system requirements of the NPPE. If temperature is not critical, 

the cold seawater will be pumped through a liquidlliquid heat exchanger cooling a treated 

fresh-water closed loop. Use of a closed loop eliminates many corrosion issues, and the 

treated water loop can operate at any design pressure. 

Power estimates for pumping the water for the VCS gas coolers are discussed in Section 6.3.7. Seawater at 6.1°C will be used in the precooler and possibly in the first intercooler device. A treated-water closed loop is planned for the remaining intercoolers and 

oil cooler. The first-cut cost estimate is based on 25 kg/s (conservative) of 6.1°C seawater flow for the precooler and 25 kg/s of fresh water for the closed loop, cooled by 

25 kg/s of 11.3"C seawater. Estimates of cost are $10,000 for three water pumps, 

$10,000 for piping (ROM), and $5,000 for the liquidlliquid heat exchangers. 

Detailed design and siting of these piping systems remains to be completed. No design 

work has been performed on the turbine oil cooler or control air compressor water loops, 

but pumping power is expected to be less than 0.5 kWe. The requirements may be integrated with the main treated-water closed-loop system at little additional cost. 
Other needs for water may be gland-sealing water for the turbine shaft seals and the 

main seawater pumps. Flow rates are expected to be low, and design of the system will 

be included in the seawater systems task. 

6.6.3 Desalinated-Water System 

The desalinated-water storage and distribution system will collect, store, and distribute 

the condensate from the surface condenser. Pumping power for these units is not part of 

the determination of net power, and the design is expected to be simple. Because the 

design is not critical to operation, little work has been completed. 

The surface condenser design is part of the heat exchanger task. After maximum net 

power experiments, up to 10% of the total steam flow rate will be routed to the surface 

condenser, resulting in a production rate of approximately 0.32 kg/s (5.0 gallmin) of 

desalinated water. The water will be pumped to a standard tank with a capacity of 

approximately 65 gal (250 L). A water fountain will be placed in the facility's visitors' 

area. Additional water taps may be made available. A simple treatment system may be 

required to ensure that potability is maintained in the storage and distribution system. 

Cost is expected to be less than $2,000. 

6.6.4 Air-Conditioning 

Building air-conditioning and ventilation system design is included in the structures task. 

Conventional systems and designs are expected, including a vapor-compression chilledwater air-conditioning system. However, effluent cold seawater at 11.3OC can also be 

used as an alternative chilled-water heat sink. This option will be carried into the 

preliminary design, but a vapor-compression system will be used for the purpose of cost 

estimation. 

2 2 The cooling load of 10 to? (35 kWe)2was estimated using 150 ft /ton (4 m /kW,) and a 

building area of 1500 ft (140 m 1. Typical costs for a vapor-compression airconditioning system using chilled water are $1,00O/ton ($280/kwe). The total airconditioning system cost estimate is $10,000. 

6.6.5 Other Miscellaneous Items 

Other miscellaneous pieces of hardware will be required to ensure the HMTEAINPPE 

operates smoothly. These small items, which do not clearly fit within the scope of the 

main component tasks, are included as part of the ancillary systems. We do not expect 

these to affect net power. Several needs identified are 

a Tool air compressor 

a Voice communication system 

a Lighting system for viewing vacuum vessel internal elements 

a Public relations materials (models, displays, etc.). 

A detailed cost estimate for these items was not attempted. A $100,000 allowance is 

included in the ancillary systems task cost estimate for these items. 

6.7 Summary 

The conceptual design process for the ancillary systems task addressed key items that 

affected the overall schedule, performance, and cost. The largest power consumer, the 
Table 6-9. Power Consumption Estimate for Ancillary 

Systems 

Power Range 

Subsystem (kW,) 

Vacuum Compression System 

Compressors 26.0 - 32.0 

Gas condenser/cooling-water pumping 5.5 - 7.0 

Lubrication pumping 0.5 - 0.5 

Oil-Cooling Water Pumping 

Turbine 0.5 - 0.5 

Vacuum compressors 0.5 - 0.5 

Instrumentation/DAS Control 

Total 

Table 6-10. Cost Estimate for Ancillary Systems 

-- - 

Subsystem 

Cost 

Estimate 

Vacuum Compression System 

Preliminary and final design 

Gas condenser/coolers 

Ancillary Electric System 

Instrument and Controls 

Computer hardware 

Instrumentation 

Miscellaneous 

Cooling water 

Desalinated water 

Building air-conditioning 

Other 

Total $988,000 
Table 6-1 1. Summary Schedule for Ancillary Systems 

Subsystem Schedule 

Vacuum System 

Preliminary contract 

Precooler/intercooler design input due 

Final design and construction contract 

Ready for HMTEA tests 

Ancillary Electrical System 

Preliminary design complete 

Final design complete 

Purchase and installation complete 

Controls and Instrumentation 

Dynamics analysis contract 

Final design contract 

Purchase and installation complete 

Ready for HMTEA tests 

Miscellaneous Systems 

Preliminary design complete 

Final design complete 

Purchase and installation complete 

vacuum system, will require engineering development to achieve desired performance, 

but manufacturers' bids and projections associated with the preliminary design indicate 

that the goal of 40 kW, can be met. This section summarizes the power consumption 

projections important In determining net power (Table 6-9), provides an overall cost 

estimate (Table 6-10), provides a revised summary schedule at the time of writing 

(Table 6-1 l), and restates the largest remaining uncertainties. 

The ancillary system costs were mentioned in the previous sections. Amounts are for 

specific hardware design contracts, hardware purchases, and installation. In-program 

labor, outside review boards, labor for system checkout, and contingencies are not 

included. Major uncertainties and assumptions regarding the ancillary systems are 

An overall compressor stage efficiency of 60% (manufacturers will address in preliminary design) 

Performance of gas condenser/coolers (researchers will analytically and experimentally address this during preliminary design) 

VCS precooler inlet conditions 

Cooling-water pumping power (researchers will fully examine flow rates, barometric 

levels, etc., in preliminary design). 

Although uncertainty remains, conceptual design investigations have not identified any 

major problems, and the potential for reaching the NPPE goals has been clearly illustrated. The schedule shows HMTEA tests starting by July 1991, which is compatible with 

the overall NPPE being ready by April 1993. 
CHAPTER 7.0 

CONCLUDING REMARKS 

This report describes the conceptual design of an experimental system developed for 

investigating heat- and mass-transfer processes and assessing the viability of open-cycle 

ocean thermal energy conversion (OC-OTEC). 

The experiment will be developed in two stages, the HMTEA and the NPPE. The goal for 

the HMTEA is to test heat exchangers to verify performance at the scale needed for the 

NPPE. The goal for the NPPE is to experimentally verify the feasibility of OC-OTEC by 

producing a minimum of 40 kWe of net power at reduced technical risk. This will be 

achieved by installing a turbine and testing the system's power-generating performance. 

The design effort was directed to meet the goals of both the HMTEA and the NPPE and 

minimize development of duplicate hardware. The choices for the design resource water 

flow rates were made consistent with the availability of cold and warm seawater as a 

result of the seawater systems upgrade carried out in 1987 in a cooperative effort among 

DOE, the State of Hawaii, and PICHTR. 

The presented design resulted from detailed considerations of pros and cons of a variety 

of choices for the system configuration. The choices were evaluated based on their 

projected performance, degree of technical risk, required schedule, and cost. 

The conceptual design as described in this report represents a viable practical arrangement for the system which poses no technical difficulties in meeting the research objectives and goals set forth. The estimated cost for the future phase of the design and the 

development of the HMTEAINPPE is consistent with the projected future program funding levels. 

The HMTEA and NPPE were designed cooperatively by PICHTR, ANL, and SERI under 

the guidance of the U.S. Department of Energy (DOE). 

The experiment will be located at the DOE'S Seacoast Test Facility (STF) at the Natural 

Energy Laboratory of Hawaii (NELH), Kailua-Kona, Hawaii. 
CHAPTER 8.0 

REFERENCES 

8.1 Open Literature 

Bharathan, D., and T. R. Penney, Jan. 1983, Flash Evaporation from Falling Turbulent 

Jets, SERI/TP-252-1853, Golden, CO: Solar Energy Research Institute. 

Bharathan, D., B. K. Parsons, and J. A. Althof, Oct. 1988, Direct-Contact Condensers for 

Open-Cycle OTEC Applications: Model Validation with Fresh-Water Experiments for 

Structured - Packines. SERIITR-252-3108. Golden, CO: Solar Ener~v Research 

Institute. 

Block, D. L., and J. A. Valenzuela, May 1985, Thermoeconomic Optimization of OCOTEC Electricity - and Water Production Plants, SERI/STR-251-2603, Golden, CO: 

Solar Energy Research Institute. 

Crane Co., 1986, Flow of Fluids Through Valves, Fittings, and Pipe, Technical Paper 

No. 410, New York, NY: Crane Co. 

Daniel, T. H., April 1989, "New Seawater Delivery Systems at the Natural Energy 

Laboratory of Hawaii," Proceedings of the Eleventh Annual ASME Solar Energy 

Conference, San Diego, CA, April 2-5, New York, NY: ASME. 

Dillingham Construction Pacific Ltd., 1987a, Solar Energy Research Institute Open-Cycle 

OTEC: Heat- and Mass-Transfer Scoping Test Apparatus, Subcontract No. 

HX-6-05107-1, Honolulu, HI: Dillingham Construction Pacific, Ltd. 

Dillingham Construction Pacific Ltd., 1987b, Open-Cycle OTEC: Heat- and MassTransfer Experimental Apparatus, final report on Contract HX-6-05107-1 for the - .. . Solar Energy Research Institute. 

Fong, K.B.K., S. M. Masutani, and L. C. Neill, July 1989, Preliminary Design of a NPPE 

Turbine based on Existing Hardware, Pacific 1ntF ~tional Center for High 

Technology Research, prepared under subcontract to tne Solar Energy Research 

Institute, Contract XX-8-07206-1. 

High Technology Development Corporation, Aug. 1985, Final Environmental Impact 

Statement, Development Plan for the Hawaii Ocean Science and Technology Park and 

Expansion of the Natural Energy Laboratory of Hawaii, prepared by The Traverse 

Group, Inc., Honolulu, HI. 

Kline, S. J., and F. A. McClintock, 1953, "Describing Uncertainties in Single-Sample 

Experiments," Mechanical Engineering, Vol. 75, No. 1, pp. 3-8. 

Lewis, L. F., L. Trimble, and J. Bowers, Sept. 1987, Open-Cycle OTEC Seawater 

Experiments in Hawaii, Proceedings, OCEANS-87, Marine Technology Soc., 

Washington, D.C., pp. 397-402. 
Lewis, L. F., J. Van Ryzin, and L. Vega, 1988, "Steep-Slope Seawater Supply Pipeline," 

Proceedings of the American society of civil Engineers, 2 1 st- ' ~nternational 

Conference on Coastal Engineering, Costa Del Malaga, Spain, 1988, New York, NY: 

Am. Soc. Civil Engineers. 

Link, H., and 8. Parsons, 1986, Potential of Proposed Open-Cycle OTEC Experiments to 

Achieve Net Power, Oceans '86 Conference Record, Washington, D.C.; Sept. 23-25, 

New York: Institute of Electrical and Electronics Engineers, Inc., pp. 207-212. 

Ludke, R. P., and P. Y. Chow, Feb. 1984, Structural Aspects of an Advanced Research 

Open-Cycle OTEC Facility, Proceedings of the Third Offshore Mechanics and Arctic 

Engineering Symposium, Vol. 2. 

Mechanical Technology Inc., May 1989, Preliminary Design and Analysis of a Single-Stage 

Steam Turbine Generator for the Open-Cycle OTEC Experimental Project, 

MTI/89TRlO, prepared under subcontract XM-8-18072 to the Solar Energy Research 

Institute, SERIIXM-8-18072-1, Latham, NY: Mechanical Technology 1ncoGorated. 

National Oceanic and Atmospheric Administration, U.S. Department of Commerce, 1989, 

Tide Table 1990, High aid Low Water Predictions, west Coast of North and. south 

America Including the Hawaiian Islands, Washington, D.C.: U.S. Dept. of Commerce. 

National Oceanic and Atmospheric Administration, U.S. Department of Commerce, 1989, 

Tide Table 1991, High and Low Water Predictions, west Coast of North and south 

America Including the Hawaiian Islands, Washington, D.C.: U.S. Dept. of Commerce. 

Natural Energy Laboratory of Hawaii, Mar. 1987, Final Supplemental Environmental 

Impact Statement, Alternative Methods of Seawater Return Flow Disposal, prepared 

by MCM Planning, Honolulu, HI. 

Ocean Data Systems Inc., 1977, OTEC Thermal Resource Report for Hawaii, Monterey, 

CA: Ocean Data Systems, Inc. 

Pacific International Center for High Technology Research (PICHTR), Dec. 1988, 

Review of the Conceptual Design for a 165-kW (Gross) OC-OTEC Plant, prepared 

under contract XX-8-07206-1 to the Solar Energy Research Institute. 

Panchal, C. B., and L. Genens, Nov. 1989, Surface Condenser Test Unit for 

HMTSTAINPPE Experiments, Argonne National Laboratory Report prepared under 

Contract W-3 1-109-Eng-38. 

Parsons, B., D. Bharathan, and J. Althof, Sept. 1985, Thermodynamic Systems Analysis of 

Open-Cycle Ocean Thermal Energy Conversion (OTEC), SERIITR-252-2234, Golden, 

CO: Solar Energy Research Institute. 

Parsons, B. K., H. F. Link, D. Bharathan, A. A. Pesaran, F. Zangrando, and C. B. Panchal, 

April 1989, Test Plan for the Heat- and Mass-Transfer Scoping Test Apparatus: Phase 

I and Phase I1 Tests, Internal Program Report, SERIIPR-253-3385, Golden, CO: Solar 

Energy Research Institute. 
Penney, T. R., J. A. Althof, D. Bharathan, and B. Parsons, Jan. 1984, Small-Scale OpenCycle OTEC Power Systems, Preliminary Concept Design, SERI/TR-252-2184, 

Golden, CO: Solar Energy Research Institute. 

Penney, T., D. Bharathan, J. Althof, and B. Parsons, 1984, Open-Cycle OTEC Research: 

Progress Summary and a Design Study, ASME Paper 84-WA/Sol-26, New York: 

American Society of Mechanical Engineers. 

Schlichting, H., 1968, Boundary Layer Theory, New York: McGraw-Hill. 

Schobeiri, T., Aug. 1989, Aero-Thermodynamic Design Study of Single-Inflow Radial and 

Single-Stage Axial Steam Turbines for OC-OTEC Net Power-Producing Experiment, 

SERIITR-253-3559, Golden, CO: Solar Energy Research Institute. 

Shelpuk, B., 1985, 165-kW Open-Cycle OTEC Experiment, SERI/TR-251-2725, Golden, 

CO: Solar Energy Research Institute, 8 pp. 

Uniform Building Code, 1988 Edition, ISSN 0896-9655, Whittier, CA: International 

Conference of Building Officials. 

U.S. Department of Energy (DOE), Dec. 1985, Federal Ocean Energy Technology 

Program, Multi-Year Program Plan, FY85-89, DOE/CH/10093-100. 

Valenzuela, J. A., T. Jasinski, W. D. Stacey, B. R. Patel, and F. Y. Dolan, June 1988, 

Design and Cost Study of Critical OC-OTEC Plant Components, ~~~1/~~~-252-3246; 

Golden, CO: Solar Energy Research Institute. 

Value Engineering Co., May 1979, Ocean Thermal Energy Conversion: Work Breakdown 

Structure, prepared under DOE contract ET-78-C-02-4931, Alexandria, VA. 

Westinghouse Electric Corp., Mar. 1979, 100 MW OTEC Alternate Power Systems, final 

report, to U.S. DOE under contract EG-77-C-05-1473, Vol. 1. 

Zangrando, F., D. Bharathan, H. Link, and C. B. Panchal, Aug. 1989, "Seawater Test 

Results of Open-Cycle Ocean Thermal Conversion Components," SERI/J-253-0395, 

Golden, CO: Solar Energy Research Institute (article accepted for publication in 

Heat Transfer Engineering Journal). 

Zangrando, F., D. Bharathan, H. J. Green, H. Link, C. B. Panchal, B. Parsons, J. Parsons, 

and A. A. Pesaran, 1990, Results of Scoping Tests for Open-Cycle OTEC components 

Operating with Seawater, SERI/TR-253-3561, Golden, CO: Solar Energy Research 

Institute. 

- 8.2 Internal Documents 

ANL letter from T. Rabas to F. Zangrando, Mar. 10, 1989, "Paper by Nisolle.ll 

ANL letter from T. Rabas to F. Zangrando, Mar. 24, 1989, "ANL Findings for the Matrix 

of Options for the Desalinated Water Production.ll 

Bharathan, D., May 18, 1989, l1Vacuum Enclosure Design Requirements/Opti~ns,~~ 

Memorandum, Golden, CO: Solar Energy Research Institute. 
De~artment of Accounting and General Services of the State of Hawaii, 1986, Proiect 8. 

"~awaii Ocean ~ciencg and Technology (HOST) Park, First ~ncrement, and seacoast 

Test Facility Upgrade, prepared by R. M. Towill Corporation. 

Doc. no. 2.1.3.6.8, H. J. Green, Apr. 25, 1989, "Pump Configurations Selected for the 

NPPE Conceptual Design." 

Doc. no. 2.1.3.6.9, H. J. Green, Apr. 25, 1989, "Structural Integrity of NPPE 

Supply/Discharge Pipes under Vacuum--Current Status." 

Doc. no. 2.1.3.6.1 1, H. J. Green, July 13, 1989, "Preliminary Seawater Systems Layout 

for the NPPE." 

Doc. no. 2.1.3.6.12, H. J. Green, Aug. 3, 1989, "Flow Rate Control for the NPPE." 

Doc. no. 2.1.3.6.14, H. J. Green, Sept. 14, 1989, "Design Conditions for the NPPE 

Seawater Systems, Update: September 14, 1989." 

Doc. no. 2.1.3.6.16, H. J. Green, Sept. 25, 1989, "Seawater Systems Pump 

Configurations." 

Doc. no. 2.1.3.6.19, H. J. Green, Aug. 22, 1989, "NPPE Seawater Systems Pump 

Configurations." 

Doc. no. 2.2.3.2, memo from F. Zangrando to A. Trenka dated Mar. 7, 1989, "NPPE heat 

exchangers subtask meeting of March 2, 1989." 

Doc. no. 2.2.3.3, letter from F. Zangrando to T. Rabas, ANL, dated Mar. 16, 1989, 

"Multiple spout results." 

Doc. no. 2.2.3.4, memo from F. Zangrando to D. Bharathan, dated Apr. 21, 1989, "NPPE 

subtask meeting of April 13, 1989." 

Doc. no. 2.2.3.5, memo from F. Zangrando to D. Bharathan dated Aug. 14, 1989, 

"Proposed NPPE evaporator configuration." 

Doc. no. 2.2.3.8, memo from F. Zangrando to G. Nix dated Oct. 26, 1989, "Meetings held 

at PICHTR on October 18-24, 1989." 

Doc. no. 2.2.3.9, memo from F. Zangrando to D. Bharathan dated Oct. 26, 1989, 

"Potential Benefits of Predeaeration in OTEC Systems and Applicability to the 

NPPE." 

DOE, 1986, Seacoast Test Facility Upgrade at Natural Energy Laboratory of Hawaii, 

cooperative agreement with the State of Hawaii. 

DOE, March 1988, Finding of No Significant Impact for Shallow Trench Disposal of 

Seawater Effluent at NELH, memorandum. 

First Working Group Meeting results, 1989, "NPPE System Development Management 

Plan," issued by the Solar Energy Research institute, March 3. 
Flug, H., Sept. 19, 1989, llComments on Proposed Concrete Vacuum Vessel,ll letter to 

ANL, Glendale, CO: United Engineers & Constructors. 

Fong, K., Oct. 1989, "Cost of Vertical Wet Pit Pump Sump Modifications." 

Green, H. J., Feb. 1989, seawater systems task presentation to the NPPE First Working 

Group Meeting. 

Masutani, S. M., Aug. 29, 1989, "NPPE Structure Options," letter to SERI, Honolulu, HI: 

Pacific Center for High Technology Research. 

OET program systems team meeting minutes, Sept. 26-27, 1989, Meridian Corporation, 

Alexandria, VA. 

Oswald, C. J., Sept. 15, 1989, "Advice Regarding Proposed Concrete Vacuum Vessel," 

letter to ANL, San Antonio, TX: Southwest Research Institute. 

Pacific International Center for High Technology Research, June 1989, "Flygt Pump 

Options for NPPE." 

Pacific International Center for High Technology Research, Sept. 1989, "Seawater Pump 

Options for NPPE." 

PICHTR letter from S. Masutani to F. Zangrando, Mar. 17, 1989, "PICHTR1s Review of 

the Matrix of Options for Desalinated Water Production." 

PICHTR letter from K. Fong to F. Zangrando, Nov. 3, 1989, llNPPE Heat Exchanger 

Conceptual Design." 

Presentation Packages, 1989, Vols. I and I1 of the NPPE First Working Group Meeting 

held at the Solar Energy Research Institute February 22-23. 

R. M. Towill Corporation, Sept. 1989, "Revised Pre-Final Bid Documents, OTEC Disposal 

System, Natural Energy Laboratory of Hawaii." 

SERI, Mar. 1989, "Results of the NPPE System Development Management Plan First 

Working Group Meeting" (February 22 and 23, 1989). 

Stallman, J. D., July 27, 1989, "Costs for Coring Concrete Walls,'' letter to SERI, 

Honolulu, HI: Concrete Coring Company. 

Stinnett, R. C., Aug. 9, 1989, "Structural Adequacy of the Concrete Ves~el,~~ letter to 

SERI, Englewood, CO: Stone and Webster Engineering Corp. 

Trenka, A. R., Feb. 26, 1987, "Meeting at DOEfHQ on February 15, 1987, Regarding 

Options for NPPE,ll SERI memorandum, Golden, CO: Solar Energy Research Institute. 
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I Document control 1 1. SERI Report NO. 1 2. NTlS Accession No. 1 3. Recrplent's Accessron No. 1 

Energy Conversion Net Power-Producing Experiment 

(OC-OTEC NPPE) 

Page TP-253-3616 I DE90000304 

4. Title and Subtitle 

Conceptual Design of an Open-Cycle Ocean Thermal 

5. Publication Date 

July 1990 

Solar Energy Research Institute 

A Division of Midwest Research Institute 

1617 Cole Boulevard 

7. ~uthoy(s) D. Bharathan, H. J. Green, H. F. Link, 

B.K. Parsons, J.M. Parsons, F. Zangrando 

9. Performing Organization Name and Address 

Golden, Colorado 80401-3393 

8. Performing Organization Rept. No. 

10. ProjecVTask/Work Unit No. nva i 

12. Sponsoring Organization Name and Address 

Solar Energy Research Institute 

A Division of Midwest Research Institute 

16. Abstract (Limit: 200 words) 

This report describes the conceptual design of an experiment to investigate heat 

13. Type of Report & Period Covered 

Technical Report 

1617 Cole Boulevard 

Golden, Colorado 80401-3393 

and mass transfer and to assess the viability of open-cycle ocean thermal energy 

conversion (.OC-OTEC) . The experiment will be developed in two stages, the Heatand Mass-Transf er Experimental Apparatus (HM.TEA) and the Net Power -Producing 

Experiment (NPPE). The goal for the HMTEA is to test heat exchangers. The goal 

for the NPPE is to experimentally verify OC-OTEC1s feasibility by installing a 

14. 

. . I turbine and testing the power-generating system. The design effort met the goals 

of both the HMTEA and the NPPE, and duplication of hardware was minimal. The I 

15. Supplementary Notes 

choices made for the design resource water flow rates are consistent with the 

availability of cold and warm seawater as a result of the seawater systems upgrade 

carried out by the U.S. Department of Energy (DOE), the State of Hawaii, and the 

Pacific International Center for High Technology Research. The choices regarding 

I configuration of the system were made based on projected performance, degree of I 

technical risk, schedule, and cost. me cost for the future phase of the design 

and the development of the IMTEAINPPE is consistent with the projected future 

program ffindiag lwels. The HMTEA and NPPE were designed cooperatively by PICHTR, 

Axgonne Nqti-onal LaGmatoq, and Solar Energy Research Institute under the 

gudance of DOE, The experiment will be located at the DOE'S Seacoast Test 

Facility at the Natural Energy Laboratory of Hawaii, Kailua-Kona, Hawaii. 

17. Document Analys-is 

a. Descriptors Open-Cycle Systems; Systems Analysis; Heat Transfer; Ocean 

Thermal Energy Conversion; Ocean Thermal Power Plants; Steam Turbines; 

Steam Condensers; Evaporators; Desalination 

b. Identifierdopen-Ended Terms 

I c. UC Categories I 

- 

18. Availability Statement 

National Technical Informat ion Service 

U. S. Department of Commerce 

5285 Port Royal Road 

Springfield, Virginia 22161 

19. No. of Pages I 

20. Price 

1 I 1 

Form No. 0069 (3-25-82)