—— ZA TR CENTRAL RESEARCH LIBRARY 3 4y5k 0361219 7 DOCUMENT COLLECTION ORNL-2605 Reactors - Power c’y,?’,? GAS COOLED;, MOLTEN SALT HEAT EXCHANGER DESIGN STUDY R. E. MacPherson CENTRAL RESEARCH LIBRARY DOCUMENT COLLECTION LIBRARY LOAN COPY DO NOT TRANSFER TO ANOTHER PERSON If you wish somaone else to see this document, send in name with document and the library will arrange a loan. OAK RIDGE NATIONAL LABORATORY operated by UNION CARBIDE CORPORATION for the U.S. ATOMIC ENERGY COMMISSION ORNL-2605 | Contract No, W-Th405-eng-26 Reactor Projects Division R. E. MacPherson Date Issued 0CT 2 81953 Osk Ridge National Laborato | Oak Ridge, Tennessee = Uns operated by ion Carbide Corporation 5 s & for the . 5, Atomic Energy Commission ¥ SYSTEMS LIBRARLES (R 3 4ySk 0361219 7 CONTENTS Abstract . . . o . o 4 e e e e e e e e e e e e e Introduction . . . & & ¢ v v ¢ ¢ ¢ 4 4t 4 e e e s e SUMBATY & & s o o o o o o o o o o o o o o s o o o o Design Considerations . . . . & v 4 o 4 o o o » « 1. Heat Exchanger Configuration . . . . . . . 2., Finned Tubing . . . . . v v v ¢« v v o o o » 5. Tubing SizZe . ¢ v v o 4 v ¢ o o & « o o o o ho Gemeral . . . . . v v 4 e e e e e e e e e DIBCUSSION « v 4 & 4 v 4 4 6 4 v v e e e e e e e e 1. Countercurrent, Cross Flow Heat Exchangers 2. Countercurrent Flow Heat Exchanger . . . . Conclusions. + v v v 4 4 ¢ v v 4 c o o o o o o o o » Method of Calculation . . . . . & & ¢ ¢ o o o o » » Nomenclature . o o o ¢« v ¢ ¢ &« ¢ o o o o o o o o o o Bibli Ogmphy L] L] L] L] LJ L ] [ ] [ ] [ ] [ ] L] . L] [} L] [} a [ [ ] L Fig. 1 Fig. 2 Fig. 3 Fig. b Fig. 5 Fig. 6 Fige T Fig. 8 Fig. 9 Fig.l0 Fig.ll Fig.l2 Fig.1l3 Fig.lh Fig.l5 i - 3 - LIST OF FIGURES Ammular Heat Exchanger Geometry . Design Study Heat Exchanger Geometry. . . . . Annular Heat Exchanger Design Parameters. . . . . . Tubing Size Optimization. . . . . « « « « . .« . Design Parameters for a Four-Pass Heat Exchanger Using Helium at 300 psig with an 8500F Inlet Temperature . . . . e & s & o 4 e s e e . Design Parameters for a Three-Pass Heat Exchanger Using Helium at 300 psig with an 8500F Inlet Temperature . . . . . » e e e s e e e e e e Design Parameters for a Three-Pass Heat Exchanger Using Helium at 150 psig with an 850°F Inlet Temperature . . . . . e . e s s e s e s . Design Parameters for a Four-Pass Heat Exchanger Using Helium at 300 psig with a TOOCF Inlet Temperatule « ¢« « o« o ¢ o o o o s o o o s o o o o Design Parameters for a Four-Pass Heat Exchanger Using Hydrogen at 300 psig with an 8500F Inlet Temperature . . . . . e e o a s e e e e Design Parameters for a Four-Pass Heat Exchanger Using Steam at 300 psig with an 8500F Inlet Temperature . « . ¢ « o o o o o o s o o o o Design Parameters for a Longitudinal Flow Heat Exchanger Using Helium at 300 psig with an 850°F Inlet Temperature . . . . . . . Heat Exchanger Container Dimensions as a Function’ of Annual Operating Cost per Heat Exchanger . . . Annual Operating Cost per Heat Exchanger as a Function of Total Blower Power Investment . . . . . Effect of Varying Allowable Salt Pressure Drop and Uranium Enrichment on Annual Operating Cost per Heat Exchanger . . . . . « ¢« ¢ « « « « & Salt Volume in Return Bends of Serpentine Fuel Tubes for a Four-Pass Heat Exchanger. . . . . . . . Page No. 17 18 19 20 21 22 2> 26 28 29 37 ABSTRACT One of the major problems in the economic evaluation of the application of forced circulation, gas cooling to high temperature, molten salt power reactor systems is the definition of the required heat transfer equipment, its size and operating cost. A design study of the. salt-to-gas heat ex- changers for such a gas-cooled system has recently been completed, and the results are reported, Helium, hydrogen and steam are considered as coolants. The effects of varying heat exchanger tubing size, coolant inlet temperature, coolant pres- sure level, allowable salt pressure -drop and uranium enrichment of the molten salt are demonstrated. The relationship between heat exchanger dimensions, fuel inventory and blower power requirements is presented in graphical form for the most pertinent cases. Comparisons are made of annual operating costs and heat exchanger overall size as a function of coolant type and ofierating conditions. Hydrogen is.:shown to be the most effective of the coolants considered, with steam and helium being roughly comparable. Assuming other conditions to be equal, helium can be made competitive with hydrogen by operating with a 50 - 60% higher helium temperature gradient through the heat exchanger. Optimum heat exchanger geometries based on gas blower power costs and en- riched fuel inventory charges require a total blower power investment of approximately 0.5% of the plant gross electrical output. However, substantial reductions in heat exchanger size can be realized by going to higher blower power investment levels. Introduction Since the early phases of design evaluation on a molten salt power reactor system, it has been desired to investigate the problems associlated with the use of gas as the primary coolant. As a step in this direction, a design study has been carried out to define the salt-to-gas primary heat exchanger which would be required in such a system. The study concerns a reactor having a thermal output of 640 megawatts (10% generated in blanket and removed through blanket cooling system) and a gross electrical output of 275 megawatts. Consideration has been given to the use of helium, hydrogen and steam as coolants. The re- ference design was based on the following: 4 primary heat exchangers 1/2" Inconel tubing (0.050" wall) Circumferential Inconel fins Helium coolant - 623 1b/sec Inlet 850°F Outlet 1025°F Pressure 500 psig Cross flow Molten Salt (Fuel 130) 1768 1b/sec Inlet 1210°F Outlet 1075°F Four pass serpentine flow In addition to determining the relative effectiveness of the three coolants, the effects of varying tube size, coolant inlet temperature, coolant pressure level, salt pressure drop and uranium enrichment were investigated., The re- sults allow a direct comparison of a gas cooled primary heat exchanger in a molten salt power reactor system with previously calculated liquid cooled heat exchangers using other salts or liquid metals as primary coolant. Since this i comparison is only one step in the overall economic evaluation required to de- termine an optimum heat transfer system, no conclusions are drawn in this report ‘ as to the desirability of adopting the gas cooling cycle for the Molten Salt Power Reactor. o Summary A design study has been completed covering the application of gas as the primary coolant in a molten salt power reactor system. The use of helium, hydrogen and steam was investigated along with the effect of gas pressure level, gas inlet temperature level and tubing size, The basic heat exchanger geometry studied was a cross, countercurrent flow arrangement with the molten salt (Mixture 130 - 62 mol % LiF, 37 mol % BeF,, 1 mol % UFh) meking four serpentine passes across the ges stream (see Fig. 2). One-half inch Inconel tubing with circumferential Inconel fins was used in all the final heat exchanger calculations. Consideration was given in the initial phase of the study to other tube sizes, and the standard size chosen is felt to approach the optimum, HEeat exchanger optimization has been based on three criteria: first, fuel inventory in heat exchanger tubesz, return bends and headers; second, gas blower power requirements; and third, minimum heat exchanger container di- mensions. Fuel inventory was evaluated at $l535/ft3/year and electiical power at 9 mills/kwh, Results of the study (Figs. 12 and 14) have shown that, at a given heat exchanger gas inlet temperature with the outlet temperature fixed, the use of hydrogen results in a smaller unit with a lower fuel inventory and power requirement than either helium or steam. However, the hazards of large scale hydrogen usage must be balanced against these obvious advantages. The optimum helium and steam heat exchanger are approximately the same in dollars invested in fuel and power but differ georetrically in that the steam unit is larger in diameter and shorter in length, Since a greater premium is attached to diameter in the construction of the required heat exchanger conteinment pressure vessel, the steam unit is judged siightly inferior to the helium unit dimensionally. However, there are lmportant incentives for the use of steanm cooling. The criginal cost of the steam inventory is negligible and the containment problem becomes of minor importance. In addition, standard and well developed auxiliary components can be used throughout a steam system. On this basis, it is con- cludec that the application of steam to the gas cooling cycle would have eco- nomic advantages over helium, Changing the coolant pressure level for a given heat exchanger configuration and heat load affects the coolant pumping power approximately as the inverse ..7_. square of the pressure ratio (i.e., doubling pressure reduces pumping power to one-fourth). Changing the coolant pressure level while maintaining a reason- . ably constant blower power input affects primarily the required face area of the heat exchanger with consequent changes in container size. Increasing the allowable salt side pressure drop increases the heat ex- changer container diameter while reducing its length., Optimization of this variable was not undertaken in the present design study. Decreasing salt enrichment by a factor of five results in a reduction of yearly heat exchanger costs by factors of 2.5-4 in the cases of primary interest. The results of a companion study on longitudinal, countercurrent flow of coolant over circumferentially finned straight tubing showed this arrange- ment to be somewhat less attractive than the comparable crossflow case. Although the heat exchanger container diameters were approximately the same, the re- quired container lengths were 30 - 40% greater. The use of such a straight tube geometry leads to thermal stress problems associated with discrete tube Plugging or flow variations from tube to tube. Design Considerations l. Heat Exchanger Configuration Two general heat exchanger configurations illustrated in Figures 1 and 2 were given detailed consideration. The first, which was ultimately rejected as the least desirable of the two, was basically an annular tubing arrangement with the salt fiow straight through and the helium in cross-countercurrent flow around disc and donut baffles. The an- nular geometry imposed no restriction on the number of helium passes across the tube bundle since the helium could be introduceé and col- lected with equal ease botk to or {rom the center of the tube bundle and to or from the outside of the bund’e., Based on correlations pre- sented in Reference 1, a three-pass arrangement wes chosen as giving a satisfactory approach to pure countercurrent heat transfer (i.e., es- sentially no correction factor to be applied to the log mean temperature difference based on the hot and cold stream inlet and outlet temperature under consideration). At the same time, as will be shown later, keeping the number of passes to a minimum results in the most compact heat ex- changer geometry. One disadvantage of this arrangement is that, since straight tubes are used running from the salt inlet to the salt outlet header, plugging ol one tube could lead to a serious differential thermal expansion corndition. To a lesser extent, flow disparities between separate tubes could cause thermal stresses to be imposed on the tubing and flow fluctuations in individual tubes could lead to strain cycling of tubing material. Since this condition could be relieved somewhat by a simple geometrical arrangement such as a right angle bend in ‘the tubing at or near either the upper or lower header, it cannot be considered as a major stumbling block to the use of this geometry. The primary obstacle appeered during the course of the study. Optimum heat exchanger geometries from a fuel inventory and overall size standpoint proved to have internal diameters which resulted in high gas velocities as shown in Fig. 3. The head losses associated with changes in coolant flow direction and ve- locivy in this geometry are hard to evaluate precisely, but it was es- timated that 2 - 3 velocity heads would be lost per pass. Since these losses approach in magnitude the losses associated with flow across the UNCLASSIFIED ORNL -LR-DWG 325586 SALT INLET ' 7 COOLANT |GAS OUTLET ' - - ///;,. ~ T st - //,/ - 3 - = = - " T /,// - | ’ /// e:_— /’ = SOMNNSNANRANY ’ < B vy - - E // i e ;i_:— //f’ / 1 \ g = s P A NS S & b ~ Y FINNED N TUBING s { N N \ \ e D N ,/7// - N s 2 Iy R N \ N \ N v \ N N N\ N g N : N N : \ ! N\ \ H N N N \ Q Nl \ \ N - SALT OUTLET HEADER .~ = = ~ | . = : J / N ’/j:. ”/;, 7z oy s E —_— ,'//lf'::- Q o = ,,//—;;/’ N~ 7 2 ey NG 2 2 > ) o ) —#—;’? - — - COOLANT = GAS INLET Fig. 1. Annular Heat Exchanger Geometry. -10- UNCLASSIFIED ORNL-LR-DWG 31210 SALT INLET SALT OUTLET v NI ,m.,/ Iy [P A\ v . . — R sy ’ ‘ I DL FINNED TUBE MATRIX COOLANT INLET Fig. 2. Design Study Heat Exchanger Geometry. -11- YT T T 2 sgasssEe: 3835 sa8acs: T : I'l T 1 1T T H T T abet 111 T H + T 1 o = t T T 3 ) | ¥) - 1 o s r » T + | B 1 H e - T 1 . | uille } I Ha ! it [ ! i + IJU a T T 18 as ane H . = 2 s i 1 b h jGsie i ; 58 ot *_ s H O ] 1381 wal pRaus FRg R BORS T i T anw — Y SEN 11 be i 4 - , T I 17 “.1..... " 1 1 + ! a T 4 | L] | I ] t t HTTH - - T ! IBE BY A : + s Raes: T r ; g ¢ [ S B T ! - P I | T M M : . £ ™~ ] H )| +t ¥ T i— g r & Esaaw 1 ! i +4 T HH 1 ! I i 8| TTE T I 1 BN t T T T + - + ! " It u i | 1 1 11 Y T + v T T 8 T wm T o i v ™ I "'ll T < b 1 B H t . 2 s : ] I + ¥ 131 t 1 i T “ S SEigssis _* h ! 188 8 8¢ 1 I T 1 + I * t f nas 99 1 1 it ¥ + - 1 4 I : t+ i AR ; : REZEESS i i 1 g A ! i T panas 10282 i HHH =it o : i H 1+ 8 T - H— PR B o i+ [ ! - 1 " + + t et 7 3 i + '8 Su ) - e =5 na NNy I+ = H T T 1 - 1 b m 1 THH 1 3 aan 1 1 M : ¥ T 1 ! yu t ] . e + 1 T ymy + H H t I f v o : * o HH T . .’ + b : " " ey S Sl “B . + ¥ . T ' ¥ O i T [ E! + ! 2 1 T b B + A } L Bt s T HH ] | ) i‘« e 1 1 JRyjusae 1 1 ? t y t ¥ b 1 t b ! ; : | : sssaMlestatas ] ! 3 b + + v H 1 M ] jssan o SROSY N M 01 T T T ++ t + I " L 18 BRSO 8 1 134+ o a 1 e H 1 b o T 7 by T e ] 1 - t B txA o 4 I AR - + t 1 T T 1 H H++ N S + f I TTe 18 o a T m; T by TT 1 iy r- T i HT A REREEEE B = ; : _ aua P BN H-+H e P 1 1 T T ¥ mm T 1 DSy ags - ] : 1 ¥ THIHY t +3 + T b ! i T 1 H + T I 588 duga T e iy & ; + Es 128s 1 . ssaan 8 fnssoess ganrEnEeS: : — p 1 s , 88 B4 pphmn sy 13s pene = - F T BEES PRBE s 4 t “ + “ I ut § . ! i ! T o I T T B p + + + L T R A BHBHS B e i m 14+ EYSEEERES s + 1 RN BREE, L1 o - 1 e + i b i e : + + T 94 ; 4 T - b 1 : + + T e 7 1 Hey o rees aei T > + ] H el m “. T i1 - H - ] T T + I t 11 1 T 1 T I8 1 ! + s} 1 T in 1 T+ : o uE : , Eosasss astass : o3 iEsssasasy o 2edas SESSE o pasasdb : 1 : 1 ] 1 : e el T ¢ InEa 11 I 1 9 + 107 %HH; : ! ! 8 gpss 3 - : Hop il : ; 4_ o8 ! I 1 T ' I 1 L e ; : +1+ — + 1 pOSaE koo by 8¢ . ; L . ’ - + 1 ! SEmansigesal t ; Tt 1 £ i 1 =S 1 - 1T 1 t T i - % { N 3 A ! - — . & - T t - : 3 + H it T Y " . t pon ; HiFis 1 4 T e T by ] t T SEEEAL spgn = EEwnres vy, : ! t ; i » ] : ! I = , 4 j ' T T 1 - } 113 1T } f H - b i3 8B 1 3 ] 1 LT : H . + t 1 } T [ s t } T 1 = & 1 1 T il n o + . T+ } } 1 I 1S B8 t e b t H H : H T T 3T 1 be b T L O N T T T Tt T t T A,_ ) T I t T t HH v a . - L] Fig. 3. Annular Heat Exchanger Design Parameters. - 12 - heat transfer surfaces, they were felt to be prohibitive, The annular bundle was therefore rejected as a suitable geometry. The second heat exchanger geometry considered was a more conventional arrangement having serpentine salt tubes with the gas flowing straight through the heat transfer matrix. In this case, it was considered neces- sary to provide a four pass arrangement to provide complete freedom for differential thermal expansion in the tubing while avoiding the float- ing header which would be required if only three passes are used. In addition, the four pass arrangement allows the salt inlet and outlet to be located close together, an arrangement which stands the best chance of reducing the amount of piping required to connect the heat exchanger and the reactor. All final optimization studies were done on the basis of the serpen- tine salt tube arrangement. A study was also made of the required heat exchanger geometry and operating costs for the case of countercurrent flow of helium over cir- cumferentially finned tubing. The finned tubing geometry was not cptimized s0 the results may not represent the best that can be done with this heat exchanger type. However, they are satisfactory as a rough tie-in with the remainder of the study. Finned Tubing Several cdesign restrictions imposed at the time the heat exchanger study was undertaken made it desirable to select a finned tubing for con- sideration that was less than the optimum from a heat transfer standpoint. Inconel tubing was chosen since data on the thermal conductivity of INOR-8, a more likely materiasl of fabrication, are currently uncertain. Homogeneous fins fabricated of Inconel were chosen to avoid ccmpletely any guestion of materials incompatability in view of the extended lifetime required of a power reactor heat exchanger. The use of nickel fins would definitely re-~ sult in a more compact heat exchanger. Copper core fins would give an im- provement over nickel but would introduce the requirement for brazing the fins to the tubing in order to protect the copper against the possibility of attack by the coolant or by impurities therein., At the present time, the introduction of the brazing requirement is considered to be undesirable - primarily from a materials compatability standpoint. On the basis of information received from the Griscom-Russell Corporation(l), a bond efficiency of 100% was assigned to the mechanical bond between the Inconel fins and tubes, The use of longitudinally finned tubing with the coolant in pure countercurrent flow was not investigated, since literature references(2’3) indicate a continuous longitudinal fin to have poor heat transfer characteristics. The use of a split longitudinal fin or pin fins might result in a competitive heat exchanger; however, these choices were not investigated since it was felt that simple mechanical bonding of such fins to the tubing could not be guaranteed to give the required degree of structural reliability. The circumferentially finned tubing configuration chosen wes ex- perimentally evaluated by Kays and London(h). The tubing dimensions listed below were scaled up from the experimental tube as indicated., Experimental Tube Design Study Tube Tube 0.D., in. 0.420 0.500 Tube I.D., in. - 0.400 Fin 0.D., in, 0.861 1.024 Fin thickness, 1n, 0.019 0.023 Fin pitch 8.72 fins/inch 7.32 fins/inch Tube pitch parallel to flow, in. 0.800 0.952 perpendicular to flow, in. 0.975 1.160 The finned tubing configuration used in the study of longitudinal coolant flow over circumferentially fimmed tubing approiimated the above. It was tested by Knudsen and Katz(5) at the University of Michigan., The actual and scaled down dimensions are as follows: Experimental Tube Design Study Tube Tube O, D., in, 0.649 0.500 Tube I. D., in. - 0.400 Fin O, D., in. 1.295 1.000 Fin thickness, in. .0255 0.0197 Fin pitch 5.85 fins/inch 7.60 fins/inch 3. - 14 - Tubing Size The majority of the design study was based on 1/2 inch tubing with & .,050 inch wall thickness. Figure U4 presents the pertinent information leading to the choice of this tubing size. It can be demonstrated that the optimum salt inventory for a given set of design conditions occurs when the unit is sized so that the salt pressure drop through the heat exchanger is at the maximum allowable value. This maximum is 36 psi, since 10% of the available 40 psi which has been customarily assigned to the heat exchanger has been utilized by entrance and exit effects. Furthermore, based on the assumption that the amount of power to be utilized in coolant circulation would be between 0.5 and 10% of the plant gross electrical output as extremes, it is possible to define the range of tube lengths and number of tubes which meet design re- quirements for a given tube size. On this basis, the lines representing the length vs number of tubes at maximum salt pressure drop for 3/4 inch, 1/2 inch and 3/8 inch tubing with 0.050 inch wall were established. The location of the lines re- presenting blower power investments of 0,5% and 10% of the plant gross electrical output demonstrate that there is a fairly narrow range of length-number of tubes combinations which will satisfy the design re- quirements, Three-fourths inch tubing was judged to be undesirable because of the excessive length requirement, although the number of tubes required for the heat exchanger was very attractive., Three-eighths inch tubing was judged somewhat unsatisfactory for the opposite reason. Although the tube length was satisfactory, the number of tubes required was Judged excessive. One-half inch tubing seemed to represent a reasonable approach to optimum, although 7/16 inch tubing might be presumed equally satisfactory, It should be pointed out that increasing the wall thickness of the heat exchanger tubing from 0.050 to 0.060 - 0.065 inch would have a negligible effect on the calculations., The total resistance of the metal wall to heat transfer normally approximated 10% of the overall resistance. [ ] oo » L . L 4 L UNCLASSIFIED ORNL-LR-DWG. 32577 \AA 2 3 a 5 & 7 8 9 1 2 3 a4 s & 7 8 9 10 10 g}/‘ == _f:;t T T I T IT == TT i Fs™ ] = ¥ i T I_'____ = e T = = 5 §E== EEai 5 5 ==t EmESSEssEnsstntceocis ) ==L = =3 R e i : SEEEIL: p= = = B —f . = aEsasTTssisil E ] S =5 = s 8 £ £2 / - s 7 H S = 2 6 — 2 | — {1 ? { - 7 " 7 o I 1 i r : s 1 ] . o ; ] 1, s / Lt - i EInMI b M S3& L i ! L i , : L il . 4 YA, 9N hann T ~ ‘¥ *L ! 3 f n i ::E 2 [ |»A Tl I Ch )y it - Fan, i 3] il n W r““ o N FAIAN i F =2 A / N N ) i 11 IBHE ' R 7 ; ! T [ . b if L E}' RN T - I 2 [AQ / . u (7] y N 48] > 4 g "_' t - i/ ] it ~NN 7 1l - y NP4 J B A | ! AD _ | | 4 L1 111 ar | [T} HEHT I 7 i i 100 200 300 500 1000 2000 3000 5000 10,000 NUMBER of TUBES Fig. 4. Tubing Size Optimization. —gl-— - 16 - General The reactor core heat load of 574 thermal megawatts was arbitrarily divided among four primary heat exchangers of 143.5 megawatts capacity each. Coolant blower efficiency was taken as 80%. Salt and helium physical properties were evaluated at their mean temperature in the heat exchanger. The pressure drop distribution in the coolant circuit was arbitrarily assigned as follows: Heat Exchanger 60% Steam Generator 30% Ducts 10% Fin efficiencies were taken from correlations presented by Gardner(B). The salt pressure drop was taken as L0 psi total, with 10% assigned to entrance and exit effects and 90% assigned to heat exchanger tube friction losses. Coolant blower power cost was evaluated at 9 mills/kwh, and a load factor of 80% was assigned to the povwer plant. Enriched fuel was assigned a yearly cost of $13§5/ft5 based on the following factors: Barren salt - $1278/ft5 1) Capitalized at 14% per year $179 u-255 - $17/gren 1) .48 Mol % UF) in fuel 2) Rental at 4%/annum $1156 $1335 In calculating coolant gas pressure drop across the tube bundle, the head loss due to flow acceleration caused by temperature and pressure change was neglected. Due to the low pressure drop and coolant tempera- ture rise, the error resulting from this assumption is well within the limits of error of the overall calculation. Discussion l. Countercurrent, Cross Flow Heat Exchangers Figures 5 to 10 present heat exchanger design study results for a glven coolant, coolant pressure level, coolant inlet temperature and number of cross flow passes. Lines of constant baffle spacing, tube bank "depth" and salt volume in the tubing are given in each case on :_- Her - PR | i - bt '] o 1 et »r ot i !% 4 = - 4 o bt H - 1 : = 1 31 T LT 1 - 5 - i fl Phma - - 1 z - - =~ T I = 2 DA VS - TG TT - TR ) S = 11 vt 11 » s = T TRIT u VBt 090 N[ WAt Tifeone l{{ T | D D S A . 1 IR EEANE) s g AF N H A s 2 LANT aml : auan u T . IREEEEERED! MHIHIIERE 111 | /! ] [ I . H 1 - H I 3 L1 TTY {' = r} 1 i BRBEE i T LT ad = 3 Fig. 5. Design Parameters for a Four-Pass Heat Exchanger Using Helium at 300 psig with an 850°F Inlet Tempera~ ture, ....[l.. -18- | LLlL Y] y Ht m- o o vz - e - 1] e 10 £ == L - -] O P o 0] X | o] Y =£ e = = ; 3 S Ry X X w L o T == "l - A Wum H o I w T=1 1 “W_ 0 %1 e o t 5 un It £ i 2 LK ] m £ : SEEN , e T @ I mm i - A b=t . 3 (1] [ e)) a e - i | - .m Pl L wv L ra 9% b > ; 1o » - » T | ¥ T 1t N aan e fl et ; £ : o g H 1 1% H H__HH n s m 'ws il » 1] O T mfl ] um i o H < ) T QO Ll nen > - 4 ! e Ll 1] EnE £l o " . - o] [ Tupt - 1] e : B! I = = HiT - : : H LhH " T T 1 1 wvi i ] i 11 O ’ o : % 3 ] T T . I i o @ mn | = et RSN ) i 8 — : 11 T T T 1! 3 ABAI ¥ l.m § Lo} = — 8 H \ w R . un e siss EEsm m oG [ | 5 ,flflwr,l ; e R _ i o el HH 4 i ¥ o T i H T T ! t c aaENNRa IR ! 4 o L IS ] ] ! ! - - o 1 i i 7] e ey DR fi L : : (] ! i P 1@.‘ + t ! . . ; ~0 - . m [ - M% 8! 1 K u - +H H - - : .mv - eREaE H L 25 SeEm: N ] t T i T T ture. ] . L - Tr - I i = B! [ T T ar T ITr R e 1Ty S [ - Il ] T LA L Ll o A B]:Icm HH - Ht+t 41 - 1 BEjag T -7 =+ N T Py H S =, L 13 ! R fa L L L —1 T 11 b b & 4 - F| ] AN i - - .w - X - 1) X 1 . = e’ " LY X - 1Y » g ._._c ] 117 ] |EA 117 jui i 117 I 1 ] - » I TOT 4 =y ek H Eh El NEY 11 " 7 11 L - 3 N 4 N h H L N L Rl 'R 7 '] i HEEERS 5 ] 1 ! - s 4 I - 10 -4 1 N D O A O 2 - il N L 4 4 E g —Hl - ; vy (N7..) 2 W WV o P t + f NAT Y o 491 [« b | SS=apss st LM — i : 0 = 1h | L 40 4 | — el e , FOL OINT N ERZ . P ABEL' d il T — T { = =d aud e - FYIT.." LY - + et et e A :j E = . e RRala i E ” !0 t: H - - N (] 3 T | @ i » & 1RRE T 1 i - = bulls wm T {lH 3 T 11 fal A 1 — T 1 Fig. 7. Design Parameters for a Three-Pass Heat Exchanger Using Helium at 150 psig with an 850°F Inlet Tempera- ture, I T - T TT7 I ] i T 11 TTTIT 1 T 3% T H i 7] SEENS 8 ] H ! EARN| TARS T I + 444 HAL R _,_q—- il 4—1‘ I -4 L. i 1 1 8 S 14 I -+ A . - A* b L 4 L4+ A 411t -~ 4 ' 1 Ly Esf T = R TgTs EEgsamgas - Saaas, EE: 3 SSEsse ] SEBE=gasads: geSSEasasstanisafaniisdtd Hiba EiE ey P P b R e TrE R T 3 hapiinsyng T 1 1 H 4+ - I H SEvgutsat MmN 8 jaREn . Tt = 14 + S = i «I— ’—4 »T—r + —+ 1 —- —— i— ’ H a E e ! T T 1444 44 ENEEAsEEEEY S 1 s — — " i {1 - NN - ] ] "r 108 Nes o W IR == - 171 | INEE [ 11 Y - H — AN 11T - e 1117 L4 1T i 3 s 1 P T P H e I U A . T 11 b Y | .Y A Y \- 11 % - b~ T 2 \ C = 11T 3 H i = RGN = HHS T . - it { 1 m D e . ; N I i 1T e "‘@ E - | 1 il | INCH fa JIEflJ]EL (L o 1 oy ” ) Y Wi r AN " L1 3 IO - Aani 1 0 Al E::.‘ flE 1 liraa = 11 - I ] . e TN L C 1 E EE“E T b "t:“ Hi T 28l RN 111 == 2 4 b EHEENDK : - LI LW L% Ll a T T ) IW | 1 T | | 1 i1 4 4L 1 Chzd L Fig. 8. Design Parameters for a Four-Pass Heat Exchanger Using Helium at 300 psig with a 700°F Inlet Tempera- ture., -21- 44 Tl 11 RE NG 1 r. V. W ALY 1T, T INE ARBEE] { r ) ] K F1 B 7 1131 TIH ER A I ) 1 b m ALXLK Eh b - 4 A Pty Y, ith an 850°F Inlet ig w Design Parameters for a Four-Pass Heat Exchanger Using Hydrogen at 300 ps Fig. 2. Temperature. ] p. T | SALT REYNOLDS R 7 - 11 I L - . 1T 4 - —_— 3 . ’ HHa -——| " be 4 i 1 - e 1T = . —+H T ! H H ITf ALY T 21t > — ) 1 - e Tl » p T T % I»-o§ 4 ] T - D R il § T 1 1 = Ll et | T 1 1 W LS AEE I & H i T . = 1 JUEE E L1 -t u 1 1 1 17 h=da2 IS 1 171 il 1 LeAr S IO ET %fi H + H x HOEE B i { ol ] u i ) N = I | I 1 - - ) A 1] 1 i ¥+t HHIRE | 1 Fig. 10. Design Parameters for a Four-Pass Heat Exchanger Using Steam at 300 psig with an 850°F Inlet Tempera- ture. _ZZ— - 23 _ & basic plot of blower power investment versus the number of tubes. The active length of each heat exchanger tube is the product of the baffle spacing times the number of passes. All heat exchangers falling along a line of constant tube bank "depth" between the number of tubes at which the fuel Reynolds Number is 3000 (3500 tubes) end the number of tubes at which the fuel pressure drop is 36 psi are satisfactory for the transfer of 143.5 megawatts from the salt to the coolant under the conditions specified. However, those units represented by the intersection of a line of constant tube bank "depth" with the line of maximum fuel pressure drop represent the optimum heat exchangers from a fuel inventory standpoint. Since tube bank "depth" is given in number of tube rows, the value mist be an integer, normally in the range of two to fifteen. Study of the figures will make clear that for a given blower pover investment there is one "best" heat exchanger geometry. As blower power is in- creased, the required number of tubes decreases until the optimum geometry for that tube bank "depth" is reached at the intersection with the maximum salt pressure drop line, If this point is inside the horizontal projection of the line representing the next higher tube bank "depth", a much larger heat exchanger will also operate at this same pover level, and further power increases require heat exchangers re- presented by points along the higher "depth" line. If the point pre- viously referred to is not inside the horizontal projection of the next higher tube bank "depth", there is a range of power values which cannot be used since no suitable heat exchanger configuration exists in this range. The values on the abscissa (Total Blower Power - % of Plant Gross Electrical Qutput) represent the propertion of 275 megawatts which is assigned to power the coolant blowers in the four primary heat exchanger circuits. The power consumption of just the four heat exchangers is 60% of the abscissa value, and the power consumption assignable to one heat exchanger is 15% of the abscissa value. Design study results are presented for the following cases: - 24 . Circuit Number of Coolant, Inlet Temp.°F Pressure,psi Passes Fig. No. Helium 850 300 4 5 Helium 850 200 3 6 Helium 850 150 5 7 Helium 700 300 i 8 Hydrogen 850 300 Y 9 Steam 850 300 L 10 2. Countercurrent flow heat exchanger Figure 11 presents the results of the study on pure longitudinal countercurrent flow over circumferentially finned tubes. The case for helium at 300 psig with an inlet temperature of 850°F is considered. In this figure, the length represents the total active length of the finned tubing and the pitch represents the tube spacing in a "delta" arrangement . Conclusions Figure 12 presents optimization curves for the various coolants and operating conditions in the form of yearly cost of fuel inventory ard blower power for one heat exchanger versus heat exchanger container length and diameter. Although an economic optimum is found for each case presented, it must be realized that the cost of heat exchanger fabrication and the effects of heat exchanger size on overall plant con- struction costs have not been considered in this presentation. By swall percentage increases in yearly operating costs above the optimum value shown in Fig. 12, sizable reductions in heat exchanger length are realized. Determination of how far one should go in this direction would be one necessary step in an overall plant economic analysis. For a given set of operating conditions, hydrogen proves to be the most atiractive coolant. If it is desired to avoid the hazards of hydrogen usage, reduction of the helium inlet temperature from 850°F tc 700°F (maintaining the outlet temperature of 1025°F constant) gives a urit smaller and cheaper to operate than is the case for hydrogen at the higher inlet temperature level. Use of a coolant inlet temperature which is lower than the freezing point of the Mixture 130 {850°F com- 1110 3 HHHRH > L4 4 LR -I-iE im i T - I - a1 - ' i sl fal ) TRl - A L s I T IR B ) IF N L4111 Fig. 11. Design Parameters for a Longitudinal Flow Heat Exchanger Using Helium at 300 psig with an 850°F Inlet Temperature. _gZ_ =26~ UNCLASSIFIED FEET LENMGTH / COST OF FUEL INVENTORY 8 BLOWER POWER FOR ONE HEAT EXCHANGER Fig. 12. Heat Exchanger Container Dimensions as a Function of Annual Operating Cost per Heat Exchanger, D Dlicates the circuit control system somewhat. To accommodate a salt - flow failure, some provision would have to be made for diversion of the coolant stream around the heat exchanger to avoid freezing the salt in the tubing. The use of steam as a coolant gas appears competitive with helium since the containment problem is minor and the gas replacement costs are negligible. The optimm steam heat exchanger is shorter but some- what larger in diameter than is the case for helium. There is es- sentially no difference in operating cost. Another strong incentive for the use of steam is the existence of a well developed technology and the availability of commercial components suited to such a system. Also presented in Figure 12 is the container dimensions for a pure countercurrent flow heat exchanger using helium with an 850°F inlet temperature in longitudinal flow over a "delta" array of circum- ferentially finned tubing. Ignoring any particular advantage this ar- rangement might possess which is outside the scope of the present study, this case does not appear as attractive as the comparable crossflow case. The container diameter is somewhat larger and the required container length is longer throughout the operating cost range of primary interest. In addition, this geometry does not possess the freedom for differential thermal expansion that is inherent in the four Pass serpentine salt tube, It should be noted that the curves of Figures 12 and 13 are not continuous as drawn (except for the countercurrent flow case in Figure 12). Since each point on the curve re?resents a tube bank "depth" in tube rows (one less or one greater than its neighbor), heat exchangers meeting design conditions and having optimm salt inventories only occur at the appropriate symbols. Figure 14 shows the effect on heat exchanger container dimensions . and on yearly operating cost for one heat exchanger unit of doubling the allowable salt side pressure drop and of cutting the uranium enrichment by a factor of five. Increasing the allowable salt side pressure drop means that the ' length of the salt flow path can be increased. Since this increases the available heat transfer surface per tube, the number of tubes can be re- duced. Figure 14 shows that the end result of this change is a heat T T A1 11111 IHHW Ji B s g ML W e S 3660 AL et ey von - - i 1- i i I: 111 ERN$ l E ERIT MENT!IAS N I i $ _E;‘ () r I i $ -IF‘ 1‘ } [ ] L ! | J _i ; | il $ 300,006 , | m%‘t il $ -280,0b i oo r d :t: LA A A NTT ’ A / | | Qr 4 i Allif | PRI 120009 ‘i PN Y . M 1 [0 4 & < I ] u = M : ; 1 ® & i : . N 7] Ok v phtil il i " ’ 1 e P T Lt - . — Anas 2 L ] = alll 4l vhy \ r H T YT prees E / o i us N - $ OO0 - = et -] L] $ H4do 2 i o= : ==Zinni $ 20,000 6 S Bt e 0 : RN ‘} i M ‘1{ :[[ Li BV mfi i} s ,!\ NS ! HT | | 005 0l 02 03 05 10 20 30 TOTAL BLOWER POWER- % PLANT GROSS ELECTRICAL OUTPUT Fig. 13. Annual Operating Cost per Heat Exchanger as a Function of Total Blower Power Investment. Y L] L . . ). [ ’ UNCLASSIFIED ORNL-LR-DWG. 32587 s it 14, Effect of Varying Allowable Salt Pressure Drop and Uranium Enrichment on Operating Cost per Heat Exchanger. Fig. Annual - 30 - exchanger container larger in diameter and shorter in length than is the case with a smaller salt side pressure drop. There is a practical limit to how far the design should be carried in this direction. When the diameter of the containment vessel becomes too large for the correspond- ing length, a change to a six salt pass geometry should be investigated. The present study was not carried this far, but the basic equations listed in Table 6 are appliceble for this purpose. Figure 1h4 also illustrates the effect of lowering uranium enrich- ment by a factor of five. In the area of interest, this reduces annual operating charges for blower power and fuel inventory to 25 - 40% of their value at the higher enrichment. The results of this study can be used to predict the heat eXchanger requirements for increased or decreased reactor power levels for the specific cases and operating conditions considered. The length of the heat exchanger is a direct function of heat load and a direct ratio can therefore be applied to this dimension, provided the change is not so great as to disproportionate the length-diameter relationship of the heat exchanger container, - %] - Method of Calculation Case - 640 thermal megawatts 576 megawatts: in reactor core 64 megawatts in blanket 275 electrical megawetts L primary:core circuit heat exchangers Helium coolant 500 psig.coolant pressure 850°F .coolant inlet' temperature 4 serpentine salt pesses Tubing - 1/2 inch Inconel, 0.050 inch wall thickness "delta" array, modified | 1.19 inch tube spacing perpendicular to flow 0.952 inch row spacing parallel to flow Fins - Inconel, mechanically bonded, circumferentially wound 1.024 inch outside diameter 0.023 inch thick 7.32 fins/inch Operating Conditions Salt Inlet Temp. 1210°F " QOutlet Temp. 1075°F ToaAr 135°F " Mean Temp. 11L43°F " Flow 1768 1b/sec Helium Inlet Temp. 850°F " Outlet Temp. 1025°F "ooAT 175°F " Mean Temp. 937°F ) Helium Pressure 300 psig Heat Load/heat exchanger k.89 x lO8 BTU/hr ) AT Log Mean 203.5°F Physical Properties Salt at 1143°F heat capacity 0.57 BTU/1b°F - 32 - viscosity 22,76 1b/ft hr ) thermal conductivity 3.5 BIU/hr £t°F ; density 122,7 1b/ft3 ) Prandtl Number 3.706 . Helium at 937°F heat capacity 1.248 BTU/1b°F viscosity 0.0865 1b/ft hr thermal conductivity 0.175 BTU/hr £t°F density 0.084 lb/ft3 specific volume 11.9 ft5/lb Prandtl Number 0.616 Inconel at 1000°F thermal conductivity 140.4 BTU/hr £4°F Salt Pressure Drop 2 APS = fs(L) vs ps . 2g 144 f = .3164 (Re)s =D + ¥y . b ¥ - 20.41 x 10° Ai ps n De NM“e’. N v - W _ b w S i o ] e i's fl(De) N ps &P = 0.0578 (La = 36 psi l. . =" )" De = 00,0333 £t (L) = .60L x 107 ()L-T° Helium Pressure Drop - (6) AP G c I o) o H s ...33.. A - 0,768 £t° £in area/ft tube 2 0,109 £t~ tube area/ft tube 0.877 £t° total area/ft tube A=0877TL.N ft2 Ac= 0.0476 L, . M ft2 £ = 0.2105 (%) (Re)co.20E§ (Re) = b The W _ 93.3 x 1o5 . D ¢ K T L .N C c L r, =41 A (7 A l=.952D ft 12 APC = 5,84 x 108 f N LE M3 ) Moo= N D &P, = 5.84 x 1o8 £ D5(L) L5 N2 6 2.8 AP, = 4.608 x 10 §L2 [_% M . L Heat Transfer q=UA, §ar, .A.T:'-AOR = (.877 LN)R $ = 792 (8) wq‘“fi““ 1.057 _c kIYB Lo Ly L Ba U hc ko AM hs AB - 3 - t¢!\r=2.645x10'5¢ Ky Ay = (178 L . MR P Ay _ .36 e hsAs hs A = (.1047 LN)R b, =k . 2.65x 107 (Re)sl'ea (Pr)sO-lL (9) B e h = L4.51 x 107 8 x-B PAp _1.85 x 1077 p 28 hs 8 hc='j 'Gc . (cp)c 2 (pr) 7 o207 () (Re)c-39 L .608 h, = 3.11 x 10 Cfi%) 1 1 (_}_IN_)608 + 2,645 x lOm5 %+ 1.85 x 10‘7 ) Nl.28 U 511 x10°\D i = ¢ . (.877 LN)R . 203.5 1 ) /LN : 605 + 2.6)4-5 X 10_3 ¢ + 1.85 x 10'7 p Nl.ég 3,11 x 107 \D , R = (L) L [ 4 1 = 178.5 (L) N 1 NP 2,645 x 2070 4 1.85 x 1071 W20 311 x 1o4¢(n p=k2 =1.75x 1072 %.522 (n )23 (D ..55... qQ = 178.5 (L) N 1.837 x 1077 (LN) 1206 | 2.645 x 1077 4 1,85 x 1077 N1.28 D +286 | .016 x 107 + .5068 ¥+ 2B D (L)N = 5.032 x 103 2) Coolant Pressure Drop &P, = 4.608 x 10° () [p]2® | Nl;Bfi' LI,J 3) Heat Transfer 86 (L)N = 5.032 x 103(__1_1\1_)' + 7.246 x 10° + .5068 W-*20 D By assumption of the number of passes, baffle spacing, L, and tube bank depth, D, the heat transfer equation can be solved for & corresponding number of tubes, Substitution of these values in the coolant pressure drop equation gives a corresponding pressure drop which can be converted to blower power consumption as follows: % of circuit pressure drop assigned to ht. ex. - 60% Volumetric flow rate through blower - 6950 ftj/sec Number of heat exchanger circuits - & Blower efficiency - 80% P%gnt Gross Electrical Output - 275 megawatts I AEM x 6950 . 100 Total Blower Power - Percent = Plant Gross Electrical Output 550 . 0.80 . 275 . 1000 . 1.341 -2 APM . 2.85 x 10 © = Power Investment - 36 - The salt pressure drop equation was used to define the number of tubes at which salt pressure drop is a maximum for a given baffle spacing, L, and number of passes. The salt Reynolds Number equation 6 (Re)s = 10.41 x 10 N defines the maximum number of tubes which can be used without going below a given Reynolds number. In all cases, 3000 was taken as the minimm desired Reynolds number. For the 1/2" tubing under consideration this defines 3470 tubes as the maxdmum number usable. Optimization curves for the various cases presented in Figuresl2, 13 and 14 were based on the parameter values taken from their respective grids at the intersection of the lines of constant tube bank "depth" with the line of maximum salt pressure.drop. This defines power investment, D, L, M and N for each case as well as the fuel volume in the tubes. Bend fuel volume was obtained from Figure 15 and header volume was calculated on the basis of a cone with a base dlemeter of 16.5 inches and a length determined by 0.1 M feet. et < -37- Fig. 15. Salt Volume in Return Bends of Serpentine Fuel Tubes for a Four-Pass Heat Exchanger. Nomenclature ..38_ A - Ac Ai AT - “u Ag (c_)- D o 0] ! | A w ! (o9} 0O O 0 1 ' s PO 0 ] fin plus tube heat transfer area/pass, £t2 coolant free flow area, ft " salt flow ares, ft2 total heat transfer area (A . R), £t2 mean tube wall area, ft2 salt side heat transfer ares, ft2 coolant specific heat, BTU/1b°F tube bank "depth", in tube rows arranged perpendicular to direction of flow salt side equivalent diameter, ft coolant Fanning (small) friction factor salt friction factor gravitational constant, ft/sec2 coolant mass velocity, 1b/sec £t° coolant heat transfer coefficient, BIU/hr £tooF salt heat transfer coefficient, BIU/hr ftoop Colburn j-factor| h (Pr)2/5 c G p Inconel thermal conductivity, BIU/hr ft2°F/in salt thermal conductivity, BIU/hr £t2°F/st tube bank depth/pass, ft baffle spacing (pass width), ft total tube length (L.R), ft number of tubes in a row perpendicular to coolant flow (M = N/D) total number of tubes coolant pressure drop/pass, lb/ft2 total coolant pressure drop (APc . R), lb/ft2 salt pressure drop, lb/in2 | . coolant Prandtl Number total heat load, BTU/hr . tube bank hydraulic radius, ft b r =h4a h c (7) 1 A -59- number of crossflow passes coolant Reynolds Number salt Reynolds Number tube wall thickness, inches iog mean temperature difference, °F overall heat transfer coefficient, BTU/hr ft coolant specific volume, ft3/lb salt velocity, ft/sec fin height, inches coolant flow rate, lb/sec salt flow rate, 1lb/sec fin thickness/2, inches fin efficiency (8) coolant viscosity, 1lb/ft sec salt viscosity, 1b/ft sec salt density, lb/ftj 2°F - 4o - Bibliography 1. Personal communication of writer with representatives of Griscom- . Russell Corporation during a visit to the Oak Ridge National Laboratory. . 2. McAdams, W. H., Heat Transmission, 3rd Edition, p. 269 ] . Norris, R. H. and Spofford, W. A., ASME, Advance Paper, New York (December, 1941). 4. Xays, W. M. and London, A. L., Compact Heat Exchangers, The National Press, 1955, p. 11k, 5. Knudsen, J. G. and Katz, D. L., Heat Transfer and Pressure Drop in Annuli, Chem. Eng. Prog. 46, 490 - 500 (1950). 6. Op.Cit., Compact Heat Exchangers, p. 2l. 7. 1Ibid, p. 3. 8. Gerdner, K. A., Efficiency of Extended Surfaces, Trans. ASME, 67, 621 - 631 (1945), 9. Amos, J. C., MacPherson, R. E., Senn, R. L., Preliminary Report of Fused Salt Mixture 130 Heat Transfer Coefficient Test, ORNL CF Memo 58-k-23, April 2, 1958, ‘4 - Table 1 Cost Comparison Data Gas Cooled Molten Salt Heat Exchanger Tubing 1/2 inch, 0.050 inch wall Inconel (1) Fins Spiral Inconel, scaled from Kays and London 8.72(C) Coolant Helium ' Coolant Pressure 300 psig Coolant Inlet Temperature 850°F Number of Passes Y Total Blower Power at Max- imum Salt Pressure Drop - .0522 217 .548 1.18 2.15 3.58 5.60 % Plant Gross Electrical Output D - Tube Bank "Depth" 2 3 L 5 6 7 8 ' L - Baffle Spacing 10.3 9.k 8.9 8.6 8.3 8.0 7.8 iy M - Tube Bank "Height" 1085 690 503 390 320 270 23 ) N - Number of Tubes 2170 2070 2010 1950 1920 1890 1850 Fuel Volume - ft.5 * Tubes 78.5 8.5 63.0 . 59,0 56.0 53.0 51.0 Bends 2.8 3.5 L,2 4.6 5.2 5.7 6.2 Headers 106.4 67:7 49,3 38.3 31.h 26.5 22.7 Total 187.7 139.7 116.5 101.9 92.6 85.2 79.9 Fuel Cost - $1355/ft3/year * 251,000 187,000 156,000 136,000 12k,000 11k,000 107,000 Blower Cost -~ 80% load factor * 2,000 9,000 23,000 51,000 92,000 153,000 240,000 9 mills/kwh - Total Annual Cost * 253,000 196,000 179,000 187,000 216,000 267,000 347,000 Minimum Container | Diameter (L + 1.5), ft 11.8 10.9 10.4 10.1 9.8 9.5 9.3 Length, ft 107.6 68.4 49.9 38.7 3L.7 26.8 22.9 * For each of four heat exchangers Table 2 Cost Comparison Data Gas Cooled Molten Salt Heat Exchanger Tubing 1/2 inch, 0.050 inch wall Inconel N Fins Spiral Inconel, scaled from Kays and London 8.72 (c)( ) Coolant Helium Coolant Pressure 300 psig Coolant Inlet Temperature T00°F Number of Passes L Total Blower Power at Max- 0.113 0.243 0.450 0,760 1.18 1.75 2.51 3.40 k.50 5.96 7.55 imumm Salt Pressure Drop - % Plant Gross Electrical Output D - Tube bank "depth" L 5 6 7 8 g 10 11 12 13 14 L - Baffle spacing 8 7.7 7.4 7.15 7.00 6.90 6.75 6.60 6.50 6.40 6.30 M - Tube bank "height" 470 366 298 250 216 189 168 150 137 125 115 N - Number of tubes 1880 1830 1790 1750 1725 1700 1675 1655 1640 1625 1610 Fuel Volume, ft3 * Tubes 52.5 49,0 46.5 44.0 2,5 1,0 39.5 38.5 37.5 37.0 36.0 Bends 3.9 4.3 4.8 5.3 5.8 6.3 6.7 7.3 7.7 8.2 8.7 Headers 46,1 35.8 29.2 2Lh.6 21.2 18.6 16.6 14.9 13.5 12.4 11.% Total 102.5 89.1 80.5 73.9 69.5 65.9 62.8 60.7 58.7 57.6 56.1 Fuel Cost - $1335/ft5/year * $1%7,000 119,000 107,000 99,000 93,000 88,000 84,000 81,000 78,000 77,000 75,000 Blower Cost - 80% loa? factor * 5,000 10,000 19,000 33,000 51,000 75,000 107,000 146,000 193,000 255,000 323,000 9 mills/kvh Total Annual Cost * $142,000 129,000 126,000 132,000 1k4L,000 163,000 191,000 227,000 271,000 332,000 398,000 Minimum Container Diameter, ft (L + 1.5) 9.5 9.2 8.9 8.65 8. 8.k 8.25 8.1 8.0 7.9 7.8 Length, ft 46,6 36.2 29.5 2k.9 21.k 18.8 16.8 15.1 13.7 12,5 ¥ For each of four heat exchangers Table 3 Cost Comparison Data Gas Cooled Molten Salt Heat Exchanger Tubing 1/2 inch, 0.050 inch wall Tnconel (1) Fins Spiral Inconel, scaled from Kays and London 8.72 (C) Coolant Hydrogen ' Coolant Pressure 500 psig Coolant Inlet Temperature 850°F Number of Passes 4 Total Blower Power at Max- 113 .238 L138 .T48 1.14 1.695 2.39 imum Salt Pressure Drop - % Plant Gross Electrical Output D - Tube bank "depth" 4 5 6 7 8 9 10 L - Baffle spacing 8.8 8.5 8.2 7.9 7.7 7.5 7.4 M - Tube bank "height" Lgs5 388 317 266 229 201 179 N - Number of tubes 1980 1940 1900 1860 1830 1810 1790 Fuel Volume, £t % o Tubes 61.0 57.0 54.0 51,5 49.5 418.0 46.5 Bends 4,1 4,6 5.1 5.6 6.1 6.7 7.2 Headers | 418.6 38.1 3.1 26.1 22,5 19.7 17.6 Total 113.7 99.7 90.2 83.2 78.1 T4k 1.3 Fuel Cost - $1335/ft37year * 152,000 133,000 120,000 111,000 104,000 99,000 95,000 Blower Cost - 80% load factor * 5,000 10,000 19,000 32,000 k49,000 73,000 102,000 9 mills/kwh | Total Annual Cost * 157,000 143,000 139,000 143,000 153,000 172,000 197,000 Minimum Container ' Diameter (L + 1.5), ft 10.3 10.0 9.7 9.4 9.2 9.0 8.9 Length, ft hg,1 38.5 31.4 26.k 22.7 19.9 17.7 * For each of four heat exchangers "g'l'{" Table L Cost Comparison Data Gas Cooled Molten Salt Heat Exchanger Tubing 1/2 inch, 0.050 inch well Inconel (%) Fins Spiral Inconel, scaled from Kays and London 8.72 (C) Coolant Steam Coolant Pressure 300 psig Coolant Inlet Temperature 850°F Total Blower Power at Max- .0148 .062 .160 J3h1 .628 1.07 1.66 2.46 3.52 4.80 6.40 immm Selt Pressure Drop - % Plant Gross Electrical Output D - Tube bank "depth" 2 3 4 5 6 7 8 9 10 11 12 L - Baffle spacing 13.0 11.9 11.2 10.7 10.3 10.0 9.7 9.4 9.2 9.0 8.9 M - Tube bank "height" 1225 783 569 Ly 362 30L 263 231 206 185 168 N - Number of tubes 2Ls0 2350 2275 2220 2170 2130 2100 2080 2060 2040 2020 Fuel Volume, ft5 * Tubes 111.1 97.5 89.2 82.8 78.0 Th. b 71.0 68.0 66.2 64.3 62.7 Bends 3.3 3.9 h.7 5.2 5.8 6.4 7.1 T.7 8.3 8.9 9.4 Headers 120.3 76.9 55.9 43,6 35.5 29.9 25.8 22,7 20.2 18.2 16.5 Total 234, 7 178.3 149.8 131.6 119.3 110.7 103.9 98.4 ok.7 Ol.h4 88.6 Fuel Cost - $1355/ft3/year * 313,000 238,000 200,000 176,000 159,000 148,000 139,000 131,000 126,000 122,000 118,000 Blower Cost - 80% load factor * 1,000 3,000 7,000 15,000 27,000 46,000 72,000 107,000 153,000 208,000 277,000 9 mills/kwh Total Annual Cost 31k,000 241,000 207,000 191,000 186,000 194,000 211,000 238,000 279,000 330,000 395,000 Minimum Container Diameter (L + 1.5), ft 14.5 13.4 12,7 12.2 11.8 11.5 11.2 10.9 10.7 10.5 10.4 Length, ft 121.5 7.7 56.4 W40 35.9 30.1 26.1 22.9 20.4 18.3 16.7 * For each of four heat exchangers - o - Table 5 Cost Comparison Data Gas Cooled Molten Salt Heat Exchanger Countercurrent Flow Tubing 1/2 inch, 0.050 inch wall Inconel Fins Spiral Inconel, scaled from Knudsen and Katz No., 5(5) Coolant Helium Coolant Pressure 500 psig Coolant Inlet Temperature . 850°F Total Blower Power at Max- .285 .530 .815 1.29 2.15 3.85 7.35 15.3 imum Salt Pressure Drop - % Plant Gross Electrical Output L - Heat transfer length; ft 40.5 38.5 37.0 35.5 33.5 31.8 30.0 28,2 L + 5 - Header spacing, ft 45,5 43,5 42.0 40.5 38.5 36.8 35.0 33.2 N - Number of tubes 2310 2250 2200 2150 2100 2040 1975 1900 P - Tube pitch (Delta), in 3.5 3.2 3.0 2.8 2.6 2.4 2.2 2.0 Fuel Volume, ft3 (3) Tubes (1) 91.7 85.4 80.6 76.0 70.5 65.5 60.3 55.0 Headers 11,2 9.8 9.2 8.6 7.8 7.2 6.6 5.8 Total ' 102.9 95.2 89.8 8L4.6 78.3 72.7 = 66.9 60.8 Fuel Cost - $1335/ft5/year () 137,000 127,000 120,000 113,000 105,000 97,000 89,000 81,000 Blower Cost - 80% load factor(3) 12,000 23,000 35,000 55,000 92,000 165,000 315,000 655,000 3) 9 mills/kwh 3 Minimum Container Diameter, ft 15.5 13.6 12.8 11.9 10.7 9.8 9.0 Length (L + 7), ft 47,5 45.5 Ly o 2,5 40.5 38.8 37.0 35. Total 149,000 150,000 155,000 168,000 197,000 262,000 bL4ok,000 736,000 (1) (2) (3) Header assumed made of two flat plates the diameter of the tube bundle and spaced apart so as to give 10 ft/sec radial velocity at periphery. Minimum container length allows 7 additional feet over that required for heat transfer, for gas inlet and header geometry. For each of four heat exchangers ....g.'.l-. Helium Helium Helium Helium Hydrogen Steam Coolant Pressure 300 psig 300 psig 500 psig 150 psig 300 psig 200 psig 3500 psig Inlet Temp, 850°F 850°F 850°F 850°F ‘fOO°F 850°F 850°F Tube Size 1/2" 3/4" 3/8" 1/2" 1/2" 1/2" 1/2" Table 6 Summary of Basic Equations Salt Pressure Drop PSI 5.99 x 10° (L) Salt Reynolds Number Nl.75 5.98 x 10l+ (L) N 3.56 x 10° (1) oy 5.99 x 10° (L) L 5.99 x 107 (L) N2 5.99 x 10° (L) Nt T2 5.99 x 10” (L) .15 10.69 x 106 N 6.57 x 106 N 15.54 x lO6 N 10.69 x 106 N 10.69 x 106 N 10.69 x 106 N 10.69 x 106 N Coolant Pressure Drop Coolant PSF Reynolds Number h.61 x 106%’_1;_”2'8 93.3 x 10° D ;] L. N 2,02 x 10 LB_'Q’E'B 93.8 x 10” D | L. N 8.0k x 10 %ETE'B 93.7 x 10° D )| T . N 922x10_(8_F2895.5x_195D L. N 1.42 x 10 _(8_3_3_2851.8 X 10° D T T.N 1.25x106%15 2.8 161 x10° D Nl. L L.N h68x10£8)_—g 253x106n L T . N - o - Coolant Pressure 300 psig 500 psig 300 psig 150 psig 300 psig 300 psig 300 psig Inlet Temp. 850°F 850°F 850°F 850°F TOO°F 850°F 850°F Tube Size 1/2" 5/4" 5/8" 1/2" 1/2" 1/2" 1/2" Table 6 - contd. Summary of Basic Equations Total Blower Power % Plant Gross Electrical Qutput 2.85 2 lO_%fiP " (L)N = 5.03 2.85 x 10'?&3M (L)N = 5.10 2.85 x lO-gAPM (L)N = 5.15 5.82 x 1o'?agM (L)N = 5.12 1.36 x 1o'3a3M (L)N = L.61 2.03 x lO-?APM (L)N = 4,90 1.48 x 10'?&PM (L)N = 6.85 Heat Transfer x 103(;g>-286+ 7.25 x 10° + .506 N0 D x 105(g_>‘286+ 4.67 x 10° + .923 N-20 D x 103(%g>'286+ 10.05 x 10° + .295 w28 D x 105(%%)'286+ 7.25 % 10° + .507 w28 ;5 D pYe 105l. H. W. Savage 14. J. A. Conlin 52. A. W. Savolainen 15. J. H. Coobs 53. H. E. Seagren 16. W. B. Cottrell 5. R. L. Senn 17. F. L. Culler 55. E. D. Shipley 18. L. B. Emlet (XK-25) 56. 0. Sisman 19. D. E. Ferguson 57. M. J. Skinner 20. J. Poster 58. A. H. Snell . 21. A. P. Frass 59. E. Storto ‘ 22. J. H. Frye, Jr. 60. J. A. Swartout 23. W. T. Furgerson' 61. E. H. Taylor - 2hk. B. L. Greenstreet 62. D. B. Trauger 25. W. R. Grimes 63. F. C. VonderLage 26. A. G. Grindell 6k. C. S. Walker 27. E. Guth 65. A. M. Weinberg 28. C. S. Harrill 66. G. D. Whitman 29. H. W. Hoffman 67. C. E. Winters 30. A. Hollaender 68. M. M. Yarosh 31. A. 5. Householder 69-70. ORNL - Y-12 Technical Library, 32. W. H. Jordan Document Reference Section 33. G. W. Keilholtz 71-90. Laboratory Records Department 34. M. T. Kelley 91. Laboratory Records, ORNL R.C. 35. B. W. Kinyon 92-93. Central Research Library 36. M. E. Lackey L. Reactor Experimental 37. J. A. Lane ) Engineering Library 38. R. S. Livingston EXTERNAL, DISTRIBUTION 95. Division of Research and Development, AEC, ORO ¢ 96-677. Given distribution as shown in TID-4500 (1kth ed.) uhder Reactors - Power category (75 copies - OTS) R