OPEN CYCLE OTEC SYSTEM WITH ~ALLING JET EVAPORATOR AND CONDENSER

A configuration for the open cycle (OC) Ocean Thernal Energy Conversion (OTEC) system is pre sented incorporating a countercurrent falling jet evaporator and a concurrent falling jet con denser. The parameters governing performance of the proposed configuration are discussed and the sizing of equipment for a 100-MWe net power out put QC OTEC plant is performed, based on recent experimental falling jet heat and mass transfer results. The perfornance of an OC OTEC plant with falling jet e•,aporator-condenser is compared with the Westinghouse conceptual design that uses an open-channel · evaporator and a surface con


INTRODUCTION
Advantages of the· open cycle Ocean Thermal Energy Conversion system, compared to the closed cycle (CC) were recognized in the early days of O't'EC development.
The UC UT!::C syscem util1zes vapor flashed from a stream of warm seawater, used as a working fluid, that actuates a very low pressure turbine.
By not using a secondary motive fluid, such as ammonia, requirements on construction materials become less stringent.· ~!oreover, rem.oval of the hu2e he>tt exchangers, which in the closed cycle transfer heat from warm seawater to a secondary fluid and from there to cold seawater with small teimperature differences, reduces capital cost and eliminates the troublesome r:iaintenance problems connected with the fouling of the heat exchanger tube surfaces.
The first successful demonstration of the technical feasibility of OTEC was performed by More efforts on OTEC development were made by the French Government until 1958, but a large production plant was not constructed because OTEC could not compete with fossil fuel energy cost at that time.
D' Arson val' s idea ofharnessing the ocean thermal gradient (1882) had· to wii t almost a century before it would be considered economically feasible.
When· renewed OTEC R&O efforts were undertaken in the United States in the wake of the Arab oil embargo of 1973, preference <l'as given to the closed cycle concept for-plants exceeding 25 MWe because the •1ery large low pressure turbine required in an open cycle plant appeared to be beyond manufacturing capability.
The current thrust of ::he DOE-s.ponsored program is devoted to implementing the closed cycle system.

THE WESTINGHOUSE QC OTEC CONCEPTUAL DESIGN
The open cycle concept, however, has not been abandoned.
A recent Westinghouse study sponsored by DOE (2-4} developed an integrated 100-l!We QC OTEC plant concept that seems to compete with similar capacity contemporary closed cycle designs. Sciubba (2) offered possible solutions in turbomachinery, packaging, and housing of the OC OTEC plant that result in ease of fabrication and substantial cost savings. Westinghouse investigated turbines with up to 140-ft tip diameters that deliver 225 rpm providing up to 100~1We net power 11nder OC OTEC conditions. rnese turbines would be fitted with fiber reinforced composite blades and a disc of hollow plate or "spiderlike" construction.
The integrated design concept (shown in Fig. 1  The spent vapor leaving the turbine rotor is condensed in a surface condenser centrally located underneath the turbine. The warm seawater leaving the evaporator and the cold seawater leaving the condenser are rejected to the ocean through a common pipe.
This circular atrangement of the evaporator and condenser interfaces nicely with the turbine.
Moreover, using the evapoi'at6t', ttirh1ne, and condenser housing walls as structural elements of the whole platform results in great savings.
The (,lestinghouse study showed a strong beneficial scale-up effect on plant cost.
A compari-· son of estimated costs of closed cycle plants with the l,lestinghouse integrated design open cycle plant indicates a break-even point for a plant capacity of about 37 HWe.
After that point the open cycle plant becomes more economical.
The calculated cost of the power module, including power system and platform, for the lOO~e CC plant f:1.tted with 90-10 Cu-Ni condenser tubes is $1825/kWe*; while the corresponding cost for the 100-!-!We OC Westinghouse power module is $1526/kWe.* The cost breakdown of the 100-.'n<e Westinghouse design (4) is rather interesting.
The size of the two major systems components is determined by an engineering compromise (Fig. 1).
The 100 Open channel flash evaporation can be quite effective at elevated temperatures when a small flashdown temperature difference is accompanied by a substantial pressure drop, which easily overcomes hydrostatic suppression of bo'iling in the liq•.lid hirer and enable'> vapor to dl..sP.n~age from the bulk of the liquid.
Therefore, it has been applied extensively· in multistage flash evaporation desalination technology where it has been studied in detail.
At the relatively low temperatures characteristic of OTEC, however, flash evaporation in channel flow is hindered by hydrostatic suppression.
Under· such conditions a channel flo,. length of about 20 m was estimated as necessary co reach a reasonable approach to thermal equilibrh•m hy the combined action of turbulent con-V!,!CL1un aw.I ,;ui:f«ce ev4p,,r.11ti.on, Thus; the large dime11i;tnn nf rhe npPn 1".hilnnel evaporator mainlv was dictated bv the method for achieving flash· evaporation, which relies on turbulent mixing to bring up eventually all the warm water close enough to the free .surface to enable vapor disengagement.
JET .. J:;YAK,QM_llQN fu' ID CONDENSATION Flash evaporation' is much more effective at low temperatures by distributing the warm seawater in falling jets, since the hydrostatic suppression of boiling is eliminated.
Heat and mass trans fer to laminar liquid jets has been adequately treated in the literature by theory and experiment (5-8).
At low noncondensible gas concentration, a condensation heat transfer coefficient exceeding 200,000 kcal/ h-m2-•c has been measured in condensing steam on a thin w;iter sheet. (R).
Although the specific mass flowrates · and jet dimensions in the experiments reported in Refs. 5, 6, and 8 are not directly applicable to OTEC jet evaporator and condenser design, the results obtained in these .works are important to the effectiveness of jet heat and mass transfer at low temperatures.
They helped dispel the long-held belief (9,10) that an important contribution to heat transfer resistance during condensation is a phase interface resistance caused by the reflection of most of the molecules hitting the surface from the vapor phase side.
By kinetic theory the phase interface conductance in condensation· is expressed by (1) ;;here hev = interface conductance f = condensation coefficient, i.e., fraction of molecules hitting the interface from the vapor side that are actually condensed 1 = heat of condensation Tv = vapor temperature Vv = vapor specific·volume ~ molecular weight R ·= gas constant A decrease in vapor temperature has a strong influence on hev because of the rapid increase in Vv associated with a temperature decrease. The variation of ·hev' according to Eq. 1, is shown in Table l for two values off. If the value of f were close to 0.04, as it was generally believed to be (9,10), Table l shows that at temperatures relevant to OTEC the interface resistance ·.rould have been the controlling factor in interphase heat and mass transfer. The ex?ected improvements to heat and mass transfer perfomance using a jet evaporator or co•\d<!:il!<!:r would he.ve be!!n very limited. The results of Refs. 5 and 8-10 have shown conclusively that the values of the coefficients of evaporation and condensation are close to l. The interface resistance to heat and mass 16D3 transfer, therefore, is relatively. unimportant even at the low O~C temperatures. At low noncondensible gas concentrations in the vapor phase the controlling resistance resides in the liquid phase and very high heat transfer rates can be expected under turbulent jet conditions. A theoretical treatment. of turbulent jet heat and mass trans.fer by Kutateladze (12) and Isachenko et al. (13) assumes the · vapor phase has no resistance to heat transfer and there is a constant eddy diffusivity· e:. This has limited practical value since it presently is difficult to obtain a good estimate fore: to use in such calculations. Besides, e: is far from being constant along the jet.
Some qualitative statements can confidently· be made about evaporation and condensation at turbulent jet How.
In the case of a pure vapor these transport phenomena are governed by turbulent mixing in the liquid phase at the temperatures .in our application.
The evaporation and .s_ondensation heat transfer coefficients he and he, respectively, averaged over the jet length therefore . increase with jet velocity at the nozzle exit section. A strong variation of the local coefficients he and he with distance fro~ nozzle exit section can he expected because or turbulence decay downstream along the jet.
The condensation heat transfer coefficient is much reduced in the presence of noncondensible gas in the vapor. This is because· a buildup of a noncondensible gas concentration gradient at the condensation interface.
No such effect is expected at the evaporation interface.
The. primary author recently performed exploratory experiments at the Department of Aeronautical· Engineering, Technion, Haifa, on heat and mass transfer between adjacent coplanar warm and cold falling water jets under temperatures and specific flowrates similar to those expected. in OTEC. The vapor flowed from the evaporation interface to the condensation interface crosscurrent in the water flow direction. Jet heat and mass transfer has found some successful large-scale application in recent years in the power generating industry. Bakay and Jaszay (13) outline the calculations in designing turbulent flat jet condensers with vapor crossflow used as part of _an indirect air cooling system in a number of power planes in eastern Europe.

FALLING JET EVAPORATOR CONFIGURATIOl'-f
Despite the very high heat transfer coefficients attainable "1th liquid jets, a crosscurrent jet configuration, geometrically similar to the one described in Ref. 13, does not lead to an acceptable .evaporator or condenser design for large open cycle application.
The vapor flow velocity required to prevent excessi'le liquid drop entrainment limits this.
For comparison consider the open channel evapo-· rater in the 100-MWe ·Westinghouse conceptual plant using a crosscurrent falling jet evaporator of co~parable capacity in which water evaporates from jets falling from radially disposed slots and flows across the evaporator periphery ~nw~rns passages that lead to Che turbine.
In the case of the Westinghouse open d1aun1:!l evaporator the cross section A 1 available to the •1apor flow leaving the warm water layer is (2) where D 1 and D 0 are the inn_er and outer diameter of the evaporator, respectively.

Figure 2. Jets Falling from Slotted Tubes of Circular Cross Section
16D4 For a crosscurrent falling · jet evaporator with the same outside diameter the cross section area A2 of the vapor leaving at the evaporator periphery is given by where Hj is the length of the falling jets.
For A 1 = ·~ the required jet length increases with plant diameter: and for the 100-MWe plant size this leads ·to R. m 19 m. This jet length represents an excess!ve loss of hydraulic head, since a rough calculation shows that the energy require.d to raise che warm water flow by l m equals about 5% of plant output. · This difficulty is circumvented in a countercurrent falling film evaporator layout as illustrated in Figs. 2-4. In this layout warm water is distributed through a number of radial manifolds to a set of circular coaxial slotted tubes disposed in a horizontal plane. Vapor flashed from the circular falling jets flows a short distance countercurrent to the falling jets and escapes into the plenum chamber above the slotted tube plane.
The cross-sectional area available to the vapor in this configuration is given by Eq. 2, just as for the open channel evaporator.

FALLING JE'l' COMDENSER CONnGURATION
Circular slotted tubes arrange<l in a coplanar concentric array, as used in the evaporator conceptual design (Fig. 4), can also be used for a falling jet condenser, In this case, radial manifolds distribute cold seawater, supplied ~y the cold water pipe, to the circular slotted The spent vapor flows downward from the turbinP through the space bet· .. ePn the . slotted. tubes concurrent with the jets and condenses by direct contact with the cold seawater •.
The effectiveness of the direct. contact condenser is very sensitive to the concentration of noncondensible gas (n. c.g,) in the condensing vapor.
The n.c.g, is entrained by the vapor towards the condensation surface.
In steady state an equilibrium sets in between the convective transportation of n.c.g. towards the liquid-vapor interface and the diffusion of n.c.g, away .from the interface driven by the n.c.g.
concentration gradient in the vapor phase.
In a study on vapor condensation by direct contact with a flowing cold water film (14) samples uE va.fJul:: toT,al:'C ,n;hdr:nm frnm pnslt.ions at ,t!,fferent distances .from the film ·surface ,and analyzed for their n.c.g. content.
To maintain constant water exit temperatures and constant

Non-Condensible Gas Distribution· Near Vapor Condensing Liquid Surface
condensation r-ates it was necessary co maintain vacuum pump suction from the test enclosure to .remove traces of n.c.g. that leaked into the system. Under steady. state conditions a strong· variation of n.c.g. concentration w with distance z was · found· in the vicinity of the condensation surface.
The experimental results (Fig. 5) followed the relation w = a exp(-bz) In an extended experiment, suction was switched ·from a port situated in the wafer exit plane 100 mm from the water surface to a location close to the surface without changing the gas mixture suction rate.
The rate of n.c.g. extraction was temporarily augmented by the increased n.c.g. concentration at the new suction port, and. the heat and mass transfer steady state was thereby disturber!. The system drifted slowly towards a new steady state at an increased condensation rate, The n.c,g, concentration decreased at a fixed location in the vapor phase.
Suction was then switched to a location close to the surface of ·che inflowing cold water, again, withou·t changing the gas mixture suction rate further improving the condensation rate, Such experiments demonstrated the importance of locating the n.c.g, extraction port at a strategic ·po-sfrion, near the entering. cold water .surface where n.c.g, concentration is highest,

To vacum manifold
Liquid falling jet Figure 6. Detail of Condenser n.c.g.

Suction Arrangement
The application of this conclusion to the removal of n.c.g. from the falling jet condenser is illustrated in Fig. 6. Segments of bent tubes are slipped on the water carrying slotted circular tubes and positioned coaxially with them by appropriate end 0-rings or grommets.

Test Cell
These ·riding tube segments are slotted on their lower s·ide, and equidistant small diameter suction tubes lead from their upper side to manifolds connected to the evacuation pump. Vapor rich in noncondensibles thus is extracted from the condenser along strips near the cold water surface at the jet entrance section where the water temperature is lowest.

THE SERI EVAPORATION-CONDENSATION TEST LOOP
A test loop now under construction at the Solar Energy Research Institute (SERI) Heat and Xass Transfer Laboratory will enable experimentation with a module of a· falling jet evaporatorcondenser in which specific flowrates and temperatures will duplicate those in a large scale OC OTEC plant (Fig. 7).

·-·"
In a vacuum enclosute a stream of warm-water is· distributed into four planar parallel falling jets by four slotted tubes fed through a manifold.
Similar cold water jets are within the vacuum enclosure parallel with the warm water jets.
Vapor flashed off the wam water jets flows from the evaporation region through the open intervals between adjacent slotted tubes towards the condensation region where it is condensed by_ direct contact with the cold water jet:s. A pump circulates warm water through a heat exchanger where the water temperature is restored by adding heat; then it goes back to the warm W"-ter distribution manifold in the test cell.

16D6
The cold water ,passes through a chiller before returning to the cold water distribution manifold.
The warm and cold water pUlllps have a 750 gmp flowrate capacity. With heater and chiller capacities of 10 6 Btu/h, evaporation-condensation experimtn~s can be performed when a temperature drop of 6°F is attained in war:u water jets with a total -flowrate of 330 gpm.
The test cell can easily be refitt ed to perform evaporation and condensation experiments with various o t her jet or film configurations. Figure 8 shows the evaporation-condensation test cell. A closeup of the evaporator slotted tubes is shown in Fig . 9. EVAPORATOR Ai.'ID CONDENSER SIZING FOR A lOO~We OC OTEC PLA!.'<T Correct sizing of a falling film evaporator and condenser requires exact data on the amount of n.c.g. to be extracted from the condenser region and on jet evaporation and condensation heat transfer coefficients within the range of s pecific flowrates and temperatures relevant to on:c . Resear::h programs have been initiated to determine such data.
A rough estimate of Che falling jet evaporator and condenser dimensions for a 100-MWe OC OTEC plant was obtained assuming that in the range of specific flowrates of 30 to 60 m 3 /h-m, jets of falling length Hj ~ 0 . 85 m produced by 1/4-in. width slots will exceed the following values of heat transfer coefficient: The nominal values of warm and cold seawateC' flowrate s and temperatures given in the Westinghouse report (3)

40.00°F
A 2.5-in. nominal size was chosen for the circular slotted tubes. The evaporator inside diameter was equated to the turbine diameter (50 m), and the distance between consecutive  The results arc shor,rn i r, T11ble 2 where w is the specific flowrate , Res is the Reynolds numbe r at the slot, n is the number of tubes, D 0 is the outer diameter of the falling Jet evaporator, and D 0 w is the outer diameter of the Westinghouse evapo r ator.
The outside diameter of the condenser is determinea by the tur b ine size, while the inside diameter is limited by the diameter of the cold water pipe .
Values of D 0 = SO m and Di 2 20 m we r-e assumed.
The condenser tubes can be spaced more <l1a<m,ely than the evapor-a t or tube s since the distahce between adjacent condensation jets is no t limited by the droplet entrainment problem.