application of matrix heat exchangers to thermomechanical exergy recovery from liquid hydrogen

Cryogenics38 (1998) 857–867 1998 Elsevier Science Ltd. All rights reserved

Printed in Great BritainPII: S0011-2275(98)00038-1 0011-2275/98/$—see front matter

Application of matrix heat exchangers tothermomechanical exergy recovery fromliquid hydrogenVikas Ahuja and Roger Green*

Department of Mechanical Engineering, University of Canterbury, Private Bag 4800,Christchurch, New Zealand

Received 4 March 1997

This paper reports the outcome of a project aimed at exploring thermomechanicalexergy recovery from liquid hydrogen.

The basis of this project was the conceptual design, development and testing of anew process for CO2 removal from air for use in alkaline fuel cells operating withhydrogen stored as a liquid, addressing simultaneously:

• thermomechanical exergy recovery from liquid hydrogen, and• its application to CO2 removal from atmospheric air.

This project was an attempt to address these issues by using the cooling availablefrom the vaporisation of liquid hydrogen and/or boil-off vapour, to remove CO2 fromthe alkaline fuel cell feed air by refrigeration purification, ie. by freezing the CO2 outof the air.

A schematic description of the process and an energy balance for refrigeration puri-fication for the CO2 removal are presented, showing that the process relies on higheffectiveness heat exchangers and water re-vaporisation. The high effectiveness heattransfer is achieved using perforated plate matrix heat exchangers. Implicit in thiswork were:

• The development of a new sizing procedure for matrix heat exchangers based onan approximate analytical solution for their performance, published recently inthis journal.

• The development of a new method for construction of perforated plate matrixheat exchangers.

• Experimental testing of matrix heat exchanger performance.• The application of matrix heat exchangers to mass transfer, and their use as revers-

ing heat exchangers.

Certain questions relating to the recent analysis published in this journal are raisedand modifications suggested.

Experimental results of heat exchanger effectiveness tests and CO2 removal testsshowed that heat exchangers of the requisite effectiveness were designed and manu-factured, and that the proposed process was successful in exergy recovery and CO2

removal 1998 Elsevier Science Ltd. All rights reserved

Keywords: matrix heat exchangers; exergy recovery; CO2 removal; liquid hydrogen;alkaline fuel cells

*Present address: Otago Polytechnic, Private Bag 1910, Dune-din, New Zealand.

Cryogenics 1998 Volume 38, Number 9 857

Application of matrix heat exchangers to thermomechanical exergy recovery: V. Ahuja and R. Green

Ntupo overall plate Ntu defined asNomenclature1

Ntupo

=1lp

+1

3f1Ntuf,1

+n

3f2Ntuf,2A heat transfer surface area, m2

Acr spacer area, m2 p plate porosityb separator width, m s separator thickness, mC mcp, J/s·K W plate width,mcp specific heat, fluid, J/kg·K ai exp(Ntuf,i)Go fluid mass velocity in perforation, kg/s·m2

b surface area per unit volume, m2/m3

G Gop, fluid mass velocity in header, kg/s·m2d plate thickness, m

h heat ransfer coefficient, W/m2·K l (ksbW)/(nsC) overall axial conductionHf fin height, m parameterkax axial conductivity, W/m·K lp (kplatedW)/(bC) plate conduction parameterkp conductivity, perforated plate in the transverse n heat capacity rate ratio

direction, W/m·K f Ntup,i/Ntuf,ikplate conductivity, plate material, W/m·Kks conductivity, spacer material, W/m·K

Subscriptsm mass flow rate, kg/sn number of plates in MHE eff effectiveNtuf,i hbd/(Gcp) f fluidNtup,i kpd/(GcpHf

2) i channel number (1 or 2)

Introduction

Hydrogen-air alkaline fuel cells (AFC’s), using hydrogenstored as a liquid, offer an attractive method for transportpower generation. Alkaline fuel cells have very fast startup times, do not require platinum for an electrocatalyst,have low operating temperatures, low structural corrosionproblems, and considerably higher power density; up to a33% weight saving compared to an acid fuel cell1. Thedevelopment of fuel cell based vehicles has focused onreformation of carbonaceous fuels such as methanol, to pro-duce hydrogen on-board, in order to be independent of theshort term availability of hydrogen. The electrolyte (KOH)in an AFC is, however, poisoned by the CO2 in the refor-med fuel and air, causing rapid deterioration in perform-ance. To date no CO2 removal system has been demon-strated that meets the criteria of cost, efficiency, weight,and volume for vehicular applications2,3. However, themain motivation for producing fuel cell based vehicles nowis the demand for zero emission vehicles and on-boardreformation does not meet this requirement. Appleby andFoulkes2 state that if pure hydrogen were readily availableand easily stored on board a vehicle, hydrogen-air alkalinefuel cells would be an obvious choice for vehicular appli-cations. Using liquid hydrogen (LH2) as the storagemedium eliminates the complexity, cost, added weight, andemissions of a reformer. There is no CO2 contaminationfrom the fuel. The air side CO2 contamination remains, butis a smaller problem.

The aim of this project was to explore the synergisticpossibilities of operating hydrogen-air alkaline fuel cellsusing hydrogen stored as a liquid, by addressing simul-taneously the issues of;

• thermomechanical exergy recovery from liquid hydro-gen, and

• its application to CO2 removal from air for hydrogen-air alkaline fuel cells.

This project was an attempt to address these issues byusing the cooling available from vaporisation of liquidhydrogen and/or boil-off vapour, to remove CO2 from the

858 Cryogenics 1998 Volume 38, Number 9

fuel cell feed air by refrigeration purification, ie. by freez-ing the CO2 out of the air.

The maximum allowable concentration of CO2 for AFCfeed air is generally taken as 10–50 ppm. Expendableabsorbers, molecular sieves, polymeric membranes withfacilitated transport and electrochemical methods have beentried, or suggested as methods for air side CO2 removal foran AFC. Details of these and several other processes, andtheir evaluation, has been presented previously by Ginerand Swette4, Piperopoulou and Bloomfield1, and Applebyand Foulkes2.

Hydrogen-air alkaline fuel cells are being explored, inseveral projects, as a power source for automobiles andlocomotives. In one of these projects, the Eureka fuel cellbus5, CO2 removal from air for the alkaline fuel cells wasbeing done using an on-board 500 kg container of soda-lime6. It would be preferable to have a regenerable on-board CO2 removal process offering minimal energy con-sumption and a substantial weight saving.

In this paper, an energy balance for refrigeration purifi-cation for the CO2 removal and a schematic description ofthe process are presented. These are used to show that theproposed process of CO2 removal by refrigeration purifi-cation using thermomechanical exergy recovery from liquidhydrogen depends on very high effectiveness heatexchangers.

Results from the experimental program used to evaluatethis CO2 removal process are given. Results of tests on CO2

removal, the use of a reversing heat exchanger to processnatural humidity, regeneration of the heat exchanger whereCO2 is removed, pressure drop, single heat exchanger andoverall system effectiveness, are shown.

The results establish that CO2 can be removed from alka-line fuel cell feed air by refrigeration purification, usingcooling available from the thermomechanical exergy ofliquid hydrogen.

Refrigeration purification calculations

Air at 1 atm must be lowered to 112 K to reduce the CO2

concentration to 10 ppm. Taking the worst case for moist-ure content in air and ambient temperature to be 100% rela-


tive humidity at 33°C, and the CO2 content in air to be384 ppm, the cooling required to lower the air temperatureto 112 K, and to condense moisture and sublime CO2 is−279.5 kJ/kg of air, the water alone requiring− 84.52 kJ/kgof air. The cooling available, for the steady flow case,assuming a LH2 storage pressure of 400 kPa, atmosphericexit pressure and an exit temperature of 280 K, is−3697.1 kJ/kg of hydrogen with a possible additional−389.7 kJ/kg of hydrogen from the heat of vaporisation.

Air is supplied to the fuel cell, in excess of the stoichio-metric amount, by an amount equal to the ratio of hydrogento oxygen utilisation factor. This ratio is usually taken as2–2.5, by volume, making the hydrogen-air mass ratio|1:85. The ratio of total cooling available to that requiredis obtained by multiplying the cooling available with themass ratio. For the case of boil-off only it is 0.154, and forvaporisation and boil-off it is 0.170.

Schematic apparatus design

The above analysis shows that only 15–17% of the coolingrequired to lower air to the calculated cold end temperature,and to condense the water and CO2, is available. This maybe further reduced by any heat leak from the surroundingsinto the apparatus. The process of refrigeration purificationis then dependent on air to air heat exchange and the re-vaporisation of the condensed water to an effectivenessgreater than at least 85% assuming the apparatus can bepartly insulated. The cooling available from the hydrogenis then used only to make up the irreversibilities in the heatexchange, the sublimation of CO2, and any heat leak.

A heat exchanger apparatus to perform CO2 removalfrom air has been designed in three main parts as shownin Figure 1:

• One reversing heat exchanger operating above the CO2

saturation temperature of| 130 K, between ambienttemperature and| 200 K, with flow switching to allowair-air heat exchange and water condensation and re-vaporisation. The concentration of water is, 1 ppm atthe normal sublimation temperature for CO2

(194.68 K).• two direct heat exchangers being cycled in the CO2 sub-

limation region, between 200 K and 112 K; and• one direct heat exchanger for hydrogen-air heat

exchange.

The reversing heat exchanger operates between ambientand| 200 K. The flow channels are exposed alternately tomoist air and then to dry air at a lower pressure after theCO2 has been removed, evaporating the previously con-densed water. The flow is switched (inlet and outletswapped) after a period dependent on how much water canbe condensed before that channel needs to be dried, andthe minimum volume flow required in each period to nullifycrossover contamination. Flow switching is achievedthrough two, four-way crossover valves controlled elec-tronically.

In the CO2 sublimation region, between 200 K and112 K, there are two direct counterflow heat exchangersoperating with no flow switching. One is on-line while theother is being dried by simply allowing it to warm up. Thevolume of solid CO2 deposited is| 0.372 ml/kgair.

For the experimental work liquid nitrogen (LN2) was


Figure 1 Schematic representation of heat exchanger appar-atus for CO2 removal. Off-line CO2 subliming heat exchangernot shown. a: ambient air inlet; b: 4 way cross-over valves,pneumatically actuated, electronically controlled, allow flowreversal in reversing heat exchanger, while there is uni-direc-tional flow in the CO2 sublimation heat exchanger; c: warm end,ambient temperature; d: reversing heat exchanger for watercondensation/evaporation, condensing and evaporatingstreams are swapped by 90° rotation of cross-over valves; e:evaporating stream; f: condensing stream; g: CO2 sublimationheat exchanger, not reversing; h: cold end, 112 K; i: bypassvalves used to control cold end temperature; j: H2 vapour fromLH2 tank; k: H2 to AFC; 1: AFC feed air with , 10 ppm CO2

used to provide the amount of cooling which would beavailable from hydrogen. A coiled tube was immersed ina LN2 bath. Cold end temperature can be controlled by abypass valve for the LN2-air heat exchanger.

The pressure drop between the air inlet and outlet, inaddition to that due to the valves and core, may beincreased to aid the re-vaporisation of water and the degree


of CO2 removal. This may be done using the bypass valvewhich also controls the cold end temperature.

Initial cool-down is envisaged being achieved by usingthe boil-off vapour resulting from initial filling of the liquidhydrogen vessel. Subsequent re-cooling of the CO2 sublim-ing heat exchangers may be done using the cold endbypass valve.

The requirement for very high effectiveness heat andmass transfer in the smallest possible volume led to theexploration of matrix heat exchangers in this role.

Matrix heat exchangers

Matrix heat exchangers (MHE’s), first described by McMa-hon et al.7 in 1949, were developed to meet the need forvery high effectiveness heat exchangers of very low vol-ume. An excellent review of matrix heat exchangers cover-ing structure, fabrication, and analytical and numericalmodels for thermal and hydraulic performance has beenpresented previously by Venkatarathnam and Sarangi8.

At low performance levels, heat exchanger effectivenessis governed by film heat transfer coefficients. For effective-ness > 90% other factors such as axial conduction, flowdistribution, and heat exchange with the surroundingsbecome important. Kroeger9 has presented a method foraccounting for performance deterioration due to axial con-duction. Kays and London10 present curves for performancedeterioration due to axial conduction for balanced flow,n= 1, andn = 0.95, based on a performance study of periodicflow heat exchangers by Bahnke and Howard11. The effectof flow maldistribution for some simple cases has beenstudied by Fleming12. For cryogenic heat exchangers thelarge temperature difference between ambient andoperating conditions is conducive to heat leak. Performancedeterioration due to heat leak has been analysed by Bar-ron13, and Chowdhury and Sarangi14.

Matrix heat exchanger performance analysis

Matrix heat exchangers were initially treated as conven-tional heat exchangers employing an extended fin surfaceon both sides. The standard hyperbolic tangent formula wasused for fin effectiveness by McMahon et al.7. Fleming15

derived a new relation for fin effectiveness, based on theassumption that the temperature difference between thefluid and the matrix, rather than the fluid temperature itself,remains constant over the length of the fin. Sarangi andBarclay16 treated the MHE as a discrete set of plate-spacerpairs, instead of being uniform in the axial direction, andfound a substantial ineffectiveness to be related to the finitenumber of plates. They assumed, however, that fin effec-tiveness was unity.

In the most recent analysis available, Venkatarathnamand Sarangi17 have derived equations governing the heattransfer in an MHE. These equations have been reduced totwo second order ordinary differential equations and fouralgebraic equations describing the energy balance and heattransfer for every plate, based on the assumptions that theaxial temperature gradient in the plate is negligible andhence the full temperature drop takes place across thespacer. They have described a numerical scheme to solvethese equations simultaneously. This analysis accounts forthe discrete plate-spacer pair sets, and non-unity fin effec-tiveness. Details of this work and a listing of the program


incorporating this numerical scheme are covered in Venkat-arathnam’s Ph.D. thesis18. Venkatarathnam19 has sub-sequently published an approximate analytical solution forthe MHE governing equations.

Venkatarathnam’s analysis was developed specificallyfor matrix heat exchangers with thermal performance of >| 90% effectiveness. He was, however, unable to manufac-ture an MHE of high enough effectiveness to perform suit-able experimental correlation. The effectiveness of theMHE used for experimental investigation reported in hisresults was only 65%. Further, the source code of the pro-gram incorporating his solution algorithm for the energybalance and heat transfer equations, as presented in his the-sis, does not run. A running version of the program in histhesis was obtained through correspondence with theauthor20. He stated that the program included in his thesisis a pared down version of a larger program and suggestedchanges to the listed source code. The modified versiongives results which concur with the graphs presented in hisvarious papers and thesis. However, the modified versionstill produces absurd results (negative values ofNtueff) ifthe axial conduction parameter is higher than some unde-fined limit. The efficacy of this solution algorithm and asso-ciated program is thus uncertain.

Sizing equation development

Venkatarathnam’s approximate analytical solution for theequations for energy balance and heat transfer, accountingfor the heat transfer coefficient, axial conduction, numberof plates, and fin effectiveness in MHE’s, for balanced flowis shown in Equation (1).

Ntueff =n(1 − a1)(1 − a2)

ln(1 − a1)(1 − a2) + 1 − a1a2 + (1 − a1)(1 − a2)Ntupo

(1)

Plate and spacer dimensions are shown inFigure 2. Venkat-arathnam has used the simplification of neglecting trans-verse resistance in the separator, by assuming the plate con-duction parameterlp = `. The axial conduction parameterl, incorporates a spacer area equivalent only to that of theseparator (bW). The parameterslp andls = nl are derivedfrom the energy balance for the separator, for the numericalsolution. Venkatarathnam has then appliedl to the approxi-mate analytical solution as the overall axial conductionparameter to account for axial conduction as done by Kro-eger9. However, Kroeger has used the overall axial conduc-tivity and total area for axial conduction to define this axial

Figure 2 Perforated plate and spacer dimensions


Figure 3 Ntueff vs. NtuD for a: Venkatarathnam, analytical withlp = `; b: this work, with finite values for lp; c: Venkatarathnam,numerical; d: Fleming and Kroeger

conduction parameter. More recently Jones and Prenger21,and Jones22 have shown analytically the relevance of con-sidering axial conduction in the matrix and the separator.Venkatarathnam’s use ofl thus represents a small pro-portion of the area over which axial conduction occurs, andthe net axial conductivity. The effect of takinglp = ` andl for the separator only, is shown inFigures 3and 4.

For this work it was necessary to predict the performanceof a matrix heat exchanger of given fixed geometry overvaried flow conditions. Venkatarathnam’s results werepresented asNtueff vs. NtuD graphs for constantf and l.An MHE of given geometry at one flow condition wouldrepresent a single point on such a graph.Figures 3and 4showNtueff vs.NtuD re-plotted for given constant geometricparameters. It is apparent that the simplification oflp = `may not be justifiable for Reynolds number flows whichare characteristic of MHE’s. Further, Venkatarathnam’sapproximate analytical and numerical solutions only showagreement providedl is calculated for the separator alone.If l is taken as the overall axial conduction parameter theapproximate analytical solution deviates considerably from

Figure 4 Ntueff vs. NtuD for a: Venkatarathnam, analytical; b:analytical, using l from (2); c: Venkatarathnam, numerical; d:Fleming and Kroeger


the numerical solution. As to whether the numerical sol-ution needs to take into account axial conduction in the heatexchanger body other than in the separator is uncertain.

For this workl has then be re-defined as

l =kaxAcr

n(d + s)C(2)

The sizing equation can then be derived from Equation (1)by re-arranging the terms and solving by substitution. Thesizing equation gives the surface area required for anyknown desired Ntu. Neglecting conduction resistance

Ntueff =n

kaxAcr

(d + s)C+

1 + e−Ntuf,i

1 − e−Ntuf,i+

2GopcpH2f

3kpd

(3)

and

A = FexpF2tanh−1(− hbd

Gcp

)G +kaxAcr

(d + s)C(4)

+2GcpH2

f

3kpdGbWHfdNtueff

If conduction resistance is included then

Ntupo =3kpdW

C(2Hf + 3S1 − p1 + pDb)

(5)

giving

A = 3expF2tanh−1(− hbd

Gcp

)G +kaxAcr

(d + s)C(6)

+GHfcp(2Hf + 3S1 − p

1 + pDb)

3kpd 4bWHfdNtueff

The variation of properties of the working fluid over thetemperature range in the heat exchanger is taken intoaccount by going through the calculations for chosenincrements of temperature.

Heat transfer and flow friction data for MHE’s

The total heat transfer resistance between the fluid streamsis made up of: convective heat transfer between the gas andthe perforated plates; and conduction along the plates, andacross the separator. The convective heat transfer processand flow friction characteristics of a matrix heat exchangerare complex. The flow cross section changes substantially,continuously, between that of the perforated plate and thatof the spacer. Consequently, the fluid undergoes alternateexpansion and contraction as it flows through theexchanger. Two stacking arrangements for the perforatedplates are possible, the perforations of adjoining platesbeing aligned or staggered. Mikulin et al.23 and Venkatar-


athnam and Sarangi24 suggest that for the staggered case,and the aligned case where spacer thickness is greater thanthe pore diameter, the boundary layer is interrupted at everyscreen maintaining developing flow. In the aligned case, ifthe spacer thickness is less than or equal to the pore diam-eter the flow through the perforations tends to becomedeveloped turbulent flow, with a secondary recirculation orstagnant zone occurring in the gaps.

Convective heat transfer. Convective heat transferoccurs at the front and back faces of the plates and in thetubular surface of the perforations. The fluid impingingcontinuously on the front face of the plates gives heat trans-fer coefficients that are an order of magnitude higher thanthose normally associated with gaseous heat transfer.

Heat transfer data and empirical correlations for perfor-ated plates have been determined by McMahon et al.7,Mikulin et al.23, Orlov et al.25, Shevyakova and Orlov26,and Hubbell and Cain27.

For this work, the equation derived by Venkatar-athnam18, combining empirical equations for heat transferfrom impinging jets on facing and lee sides of a plate andthe cylindrical surface of perforations, has been used forthe design of the perforated plates. This was done with cor-relation of the experimental work to his analytical work inmind. This equation compares well at low Reynolds num-ber with the empirical equations for perforated platederived by Hu et al.28.

Flow friction. Flow friction data and empirical corre-lations for drag coefficient for perforated plate have beendetermined by several authors. The results of McMahon etal.7, Mikulin et al.23, Shevyakova and Orlov26, and Hubbelland Cain27 all show that flow friction varies minimally withReynolds number at Re > 200, indicating that form dragpredominates.

The convective heat transfer and flow friction character-istics of perforated plate are strongly dependent on theshape and size of the perforations, porosity, plate andspacer thickness, alignment of perforations on adjoiningplates, as well as the method of manufacture. Hubbell andCain27 show different heat transfer results for perforatedplate manufactured by punching, with the perforationbreakout material facing upstream, and downstream. Thisstrong geometric dependence of heat transfer and flow fric-tion characteristics is in keeping with the lack of agreementin the results of the various researchers.

The effects of spacer thickness on the heat transfer coef-ficient and flow friction depend on the manner of stacking.For the staggered case increasing the spacer thickness topore diameter ratio can be expected to reduce the drag coef-ficient while not substantially effecting the convection heattransfer coefficient. This is backed up by the results ofMikulin et al.23, Orlov et al.25 and Shevyakova and Orlov26.For aligned holes increasing the spacer thickness to porediameter ratio may be expected to increase both the heattransfer and drag coefficient, and this is confirmed by theresults of McMahon et al.7 and Mikulin et al.23.

Matrix heat exchanger construction

Since McMahon et al.’s7 first use of punched plates andneoprene gaskets in a demountable assembly, severalmethods have been devised for MHE construction. Venkat-arathnam and Sarangi’s8 review lists several of these


methods. For this work, following the failure of attemptsto manufacture wire mesh cores in-house, a new methodfor fabricating perforated plate MHE’s was devised. Thismethod, described by Ahuja29, was based on easily andcheaply available materials and processes. It allows a metalperforated plate-plastic spacer structure, with the finalassembly requiring only a low temperature oven.

Apparatus

Heat exchanger effectiveness, at steady state, may be calcu-lated directly from inlet and outlet temperatures. At higheffectiveness however, errors in temperature readings maybe amplified in terms of Ntu. Gifford et al.30 developed analternative method to measure the effectiveness of regener-ators operating at cryogenic temperatures. In this methodthe warm fluid exiting the test regenerator flows througha liquid nitrogen (LN2) bath before entering an identicalregenerator as the cold fluid. The LN2 is used to providea constant cold end temperature and the boil-off of LN2from this bath provides a direct measure of theineffec-tivenessof the regenerator. Lins and Elkan31 have used thismethod for measuring the effectiveness of wire meshMHE’s. Venkatarathnam18 has developed an apparatus formeasuring MHE effectiveness which is similar in principle.He has introduced a by-pass which allows, only part of thefluid flow, to be diverted through the LN2 bath while therest goes through the by-pass. The relative flows in the by-pass and the LN2 bath are each controlled by a valve. Thisallows control of the cold end temperature. The two valvescan also be used to increase the pressure difference betweenthe cold and hot streams.

The test apparatus used for this project is shown sche-matically in Figure 5. On exiting the hot side channel ofthe test MHE the working fluid stream is divided into twoparts. One part passes through the by-pass valve and entersthe return (or cold side) channel of the test MHE. The otherpart is diverted through the LN2 heat exchanger valve, andthrough the liquid nitrogen heat exchanger (LN2-HX). TheLN2-HX is simply a copper tube immersed in a LN2 bath.The proportion of the fluid diverted through the LN2-HXis controlled by the by-pass and LN2-HX valves. The com-bined stream then flows through the return channel of thetest MHE.

The liquid nitrogen bath is thermally shielded by asecond liquid nitrogen bath referred to as the outer bath orguard vessel. This is done to minimise boil-off in the innerbath, due to heat leak from the surroundings. The heatexchanger, bypass and LN2-HX valves, and LN2 baths aresuspended inside, from the lid of, a stainless steel vacuumvessel. The vessel is approximately 400 mm in diameterand 600 mm high. The vessel is maintained at a pressureof 10−3 to 10−6 torr by a diffusion pump vacuum system.The vacuum provides the thermal insulation necessary toeliminate excessive heat leak from the surroundings intothe heat exchangers under test. The insulation is aided bya polished copper radiation shield immediately inside thevessel. The radiation shield is cooled by boil-off vapourfrom the guard vessel.

Instrumentation

All temperature measurements were done using K typethermocouples. The thermocouple probes inside the vac-


Figure 5 Schematic lay out of test apparatus and heat exchanger test vessel

uum vessel are in stainless steel sheaths and have groundedjunctions. The flow rate of the boil-off of LN2 was meas-ured using a pelton wheel turbine meter or a thermal massflow meter, depending on the flow rate range. The workingfluid flow rate in the MHE was measured using a variablearea flowmeter. A piezoresistive pressure transducer wasused to measure the pressure drop across the MHE. Dataacquisition is done using an I/O board interfaced with a 16channel multiplexer and amplifier board and other purposebuilt electronics.

For the experimental work, the apparatus size is basedon a 1 kW AFC system. The volume of each of the matrixheat exchangers is 1.37 dm3 including the headers. Heattransfer surface area on each side is 0.64 m2 (4000 m2/m3

for the perforated plate). Design flow rate is70 l/min(nominal) of air.

Results

The experimental programme was conducted as four separ-ate series of tests to determine:


1. Heat transfer and pressure drop performance of thematrix heat exchanger (MHE). To establish that theheat transfer effectiveness of the designed heatexchangers was greater than at least 85%, as requiredfor the CO2 removal process. This performance wasmeasured by conducting tests on a single heatexchanger.

2. Heat transfer and pressure drop performance of twoMHE’s coupled together. The CO2 removal processdesign requires two MHE’s to be connected in serieswith a cross-over valve between them. It was notknown what effect this configuration would have onthe combined performance and this had to be determ-ined experimentally.

3. The feasibility of CO2 removal from air to 10 ppm byrefrigeration purification. These tests were carried outusing bottled dry air so that the CO2 removal alonecould be measured.

4. CO2 removal from ambient humid air, and the perform-ance of the reversing heat exchanger in water conden-sation and re-vaporisation (water cycling).


Heat exchanger performance

The results of single heat exchanger effectiveness tests areshown inFigure 6. Effectiveness is plotted over a range offlow rates around the design flow rate of 70 l/min. Theresults show the effectiveness to be| 97%. Each experi-mental data point on the graph represents the mean of datarecorded for 2 to 4 separate tests. In a typical test 150 datasets were recorded over 10 to 20 minutes of steady stateoperation. The calculated effectiveness is the combinedresult from that calculated from temperature and boil-offmeasurements.

In this apparatus the use of the by-pass circuit and asso-ciated valves is the main source of heat leakage. Heat leakeffects both temperature and boil-off measurements, and isdependent on the mass flow rate of the working fluid. Sincethe effect of this heat leak can not be completely dis-tinguished from the ineffectiveness of the heat exchangersit is reflected in the minimum obtainable experimentaluncertainty which is at best6 1%. Figure 6 shows thetheoretical performance prediction based on the numericalsolution of Venkatarathnam18 and that using the modifiedanalytical solution alongside the experimental results.Within the limits of experimental uncertainty the threeresults can not be significantly distinguished. It should benoted that none of the cited literature which reports experi-mental results of MHE testing includes a quantitative state-ment of experimental uncertainty. Given the presence ofthis experimental uncertainty no justifiable conclusion onthe trends of calculated and measured effectiveness versusRe, or the particular validity of any model, can be drawnfrom this data.

Figure 7 shows the pressure drop across one side of theMHE as measured and as predicted. One predicted resultis from the empirical equation of Shevyakova and Orlov26.The other was calculated using graphs from Kays and Lon-don10 for entry and exit pressure loss coefficients for mul-tiple-circular-tube heat exchanger core10 and mean frictionfactor for the hydrodynamic entry length of a circulartube[10b]. Venkatarathnam18 has used the same graphsfrom Kays and London, for pressure drop prediction, butcontrary to the suggestion of Kays and London10 for inter-

Figure 6 Heat exchanger effectiveness vs. nominal flow rate,for single matrix heat exchanger for a: this work, modified ana-lytical; b: Venkatarathnam, numerical; c: this work, experimentalresult, confidence 99.7%; d: design point


Figure 7 Pressure drop across one channel of a matrix heatexchanger vs. nominal flow rate of nitrogen for a: experimentalmeasurement; b: predicted from Kays and London; c: predictedfrom Shevyakova and Orlov

rupted fin surfaces, he has not usedKc and Ke values forRe = `. Further, he has used theapparentfriction factorwhich already includes the entry and exit loss, rather thanthemeanfriction factor. As may be seen fromFigure 7 theresult predicted using theKc andKe values for Re= `, andthe mean friction factor are in excellent agreement with theexperimental result.

Results of effectiveness tests of the two heat exchangerscoupled together including the cross-over valves and asso-ciated pneumatic actuator are shown inFigure 8. The effec-tiveness is slightly lower,| 93%. This is due to heat leakthrough the cross-over valve between the two heatexchangers. The valve actuator was pneumatic and in theearly experiments would freeze and leak when cold. It washeated to prevent this from happening. However there is agood conduction path from the actuator through the cross-over valve and this heating adds to the total heat leak intothe process stream.

Figure 8 Combined effectiveness vs. nominal flow rate, fortwo matrix heat exchangers in series. Confidence 99.7%. a: cal-culated from boil-off 6 4.4%; b: calculated from temperatures6 3.0%; c: combined result; d: regression line for experimentalresult; e: minimum uncertainty 6 2.5%


Figure 9 CO2 concentration in outlet air and pressure differen-tial between inlet and outlet plotted over the period of a test.a: CO2 concentration in outlet air, time averaged per bottle; b:pressure differential; c: cold end outlet average temperaturedeviation from 112 K

CO2 removal process evaluation

CO2 concentration was measured using a quadrupole massspectrometer. This was intended to allow continuous moni-toring of CO2 and H2O. It was found that the concentrationof both these, for a given sample, increases over time. Thismay be because both are strongly adsorbed on surfaceswithin vacuum systems, CO2 in particular on the analyserfilament32, and this leads to a build up of CO2 and H2O inthe analyser chamber. As a result of this, CO2 concentrationhad to be measured as a differential between inlet and outletair streams and averaged over time. The period over whichthe CO2 concentration was averaged was the duration abottle of air would last during testing. This was donebecause there was some variation in the CO2 levels in dif-ferent bottles.

CO2 removal test results are shown inFigures 9and10.CO2 concentration in the outlet air is plotted over the periodof a test. Figure 9 shows that CO2 removal to below10 ppm is possible by this technique.Figure 10shows thatthe amount of CO2 remaining in the treated air is below

Figure 10 CO2 concentration in outlet stream vs. time, aver-aged results from all tests. Confidence 68.3%


20 ppm for the up to 1 1/2 hours, and below 50 ppm forup to 2–21/2 hours. The level increasing as the cold endof the heat exchanger is fouled by solid CO2 deposition.

Figure 9 also shows the pressure drop across the heatexchanger apparatus over the period of the test. It can beseen that the pressure drop increases with time. This is dueto the deposition of solid CO2 in the cold end of the heatexchanger operating in the CO2 sublimation temperaturerange. The increase in pressure drop can be used to decidethe cycle time for the cold end heat exchangers.

The pressure drop across each channel of the heatexchanger core, without any CO2 or H2O deposition, wasmeasured to be 0.95 kPa at 70 l/min (| 4 kPa for the twoheat exchangers). With a total pressure drop of 10 kPa theremaining drop is mainly due to the cross-over valves.

Cold end temperature was controlled manually and hencevaried somewhat throughout a test.Figure 9shows the tem-perature deviation from 112 K at the cold end over the per-iod of a test.

The cross-over valves and associated actuator causedseveral problems. Heat leak and pressure drop have beenmentioned earlier. The heat leak through the valve and actu-ator prevented the middle stage temperature from reaching200 K. In addition the cross-over valve itself leaked whencooled below 233 K. This was due to the differentialshrinkage of the gland packing. During the test this tem-perature would rise to about 238 K over a period of 31/2to 4 hours. The saturation concentration of water in air at233 K is 128 ppm and at 238 K is 222 ppm. It may beexpected that this amount of water would carry over intothe CO2 subliming heat exchanger and be deposited duringthe test. The measured difference between inlet and outletconcentration of water was on average 143 ppm, con-firming the carry over of water due to the temperature limitimposed by the cross-over valve.

The CO2 deposited in the cold end heat exchanger maybe expected to vaporise if the temperature at the cold endis allowed to rise. The saturation temperature for CO2 inair is 130 K. At the end of a CO2 removal and water cyclingtest, the cold end temperature was allowed to rise from itstest value of 112 K.Figure 11shows the measured pressuredrop decrease and cold end temperature rise with time. Itcan be seen that as the temperature rises from 112 K to

Figure 11 Regeneration test for CO2 subliming heatexchanger. a: cold end outlet average temperature deviation; b:pressure differential across inlet and outlet


130 K the pressure drop across the heat exchangers returnsfrom 30 kPa to the original| 10 kPa.

In addition to heat exchanger performance and the CO2

removal process, the new MHE construction was evaluated.The two MHE’s built for this project were used over aperiod of two years. They were cycled from ambient tem-perature to near liquid nitrogen temperature on at least 30occasions, and were tested in operation at cryogenic tem-peratures for approximately 120 hours. During cool-downtests the MHE’s were flooded with liquid nitrogen, and laterduring tests using air the cold end was flooded with liquidoxygen. The heat exchangers have been used at pressuresof up to 4 bar. None of these factors resulted in any leaks,any visible degradation of their construction, or change inthermal-hydraulic performance.

Conclusion

The research undertaken in this project established that CO2

can be removed from alkaline fuel cell feed air by refriger-ation purification, using the cooling available from thermo-mechanical exergy recovery from liquid hydrogen, and thatthis can be done in a regenerable manner.

The experimental work conducted showed that CO2

removal from air to below 10 ppm is possible using theproposed process. Water condensation and re-vaporisationwhich must occur as part of this process, has been demon-strated using a reversing heat exchanger. Regeneration forthe heat exchanger in which the CO2 was deposited wasachieved by allowing it to warm up through a temperaturerange of 18 K. The increase of the pressure differentialbetween inlet and outlet air streams was linked to outletCO2 concentration. This fact suggests a method for con-trolling regeneration timing. Heat transfer effectiveness wasmeasured at 93%6 2.5%. These results collectively indi-cate that the proposed CO2 removal process functioned asdesigned, with a 93% internal refrigeration recovery. Thetesting has demonstrated that even with a hydrogen-airmass ratio of 1:85 for normal alkaline fuel cell operation, itis possible to effect CO2 removal using thermomechanicalexergy recovery if high effectiveness heat exchangers areused. With improved design of the cross-over valves andactuator used with the reversing heat exchanger, it is quitefeasible that such a process could be used in vehicularapplications.

Acknowledgements

Funding for this research program was received from: Uni-versity of Canterbury Postgraduate Scholarship for V.Ahuja; New Zealand Foundation for Research, Science andTechnology, Public Good Science Fund.

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