Refueling-station costs for metal hydride storage tanks on board hydrogen fuel cell vehicles

international journal of hydrogen energy xxx (xxxx) xxx

Available online at www.sciencedirect.com

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Refueling-station costs for metal hydride storage tanks on board hydrogen fuel cell vehicles Edward D. Frank*, Amgad Elgowainy, Yusra S. Khalid, Jui-Kun Peng, Krishna Reddi Argonne National Laboratory, 9700 South Cass Avenue, Lemont, IL, 60439, United States

highlights

graphical abstract


Metal hydride storage tanks on board fuel cell vehicles reduce modeled fueling costs.
Costs were lower compared to 700bar

tanks

because

of

lower

compression costs.
Savings were insensitive to uncertainties other than station labor and tank pressure.
A key question is whether selfservice

is

possible

despite

required coolant lines.

article info

abstract

Article history:

Refueling costs account for much of the fuel cost for light-duty hydrogen fuel-cell electric

Received 22 March 2019

vehicles. We estimate cost savings for hydrogen dispensing if metal hydride (MH) storage tanks

Received in revised form

are used on board instead of 700-bar tanks. We consider a low-temperature, low-enthalpy

24 September 2019

scenario and a high-temperature, high-enthalpy scenario to bracket the design space. The

Accepted 27 September 2019

refueling costs are insensitive to most uncertainties. Uncertainties associated with the cooling

Available online xxx

duty, coolant pump pressure, heat exchanger (HX) fan, and HX operating time have little effect on cost. The largest sensitivities are to tank pressure and station labor. The cost of a full-service

Keywords:

attendant, if the refueling interconnect were to prevent self-service, is the single largest cost

Hydrogen refueling station

uncertainty. MH scenarios achieve $0.71e$0.75/kg-H2 savings by reducing compressor costs

Techno-economic analysis

without incurring the cryogenics costs associated with cold-storage alternatives. Practical

Simulation model

refueling station considerations are likely to affect the choice of the MH and tank design.

Storage

© 2019 Hydrogen Energy Publications LLC. Published by Elsevier Ltd. All rights reserved.

Fuel cell electric vehicles

Abbreviations: ACHX, air-cooled heat exchanger; CEPCI, Chemical Engineering Plant Cost Index; DOE, U.S. Department of Energy; EG, ethylene glycol; HDSAM, Hydrogen Delivery Scenario Analysis Model; HFCEV, hydrogen fuel cell electric vehicle; HX, heat exchanger; IRR, internal rate of return; MH, metal hydride; PG, propylene glycol; TEA, techno-economic analysis; VFD, variable-frequency drive; k$, thousands of US dollars. * Corresponding author. E-mail address: [email protected] (E.D. Frank). https://doi.org/10.1016/j.ijhydene.2019.09.206 0360-3199/© 2019 Hydrogen Energy Publications LLC. Published by Elsevier Ltd. All rights reserved. Please cite this article as: Frank ED et al., Refueling-station costs for metal hydride storage tanks on board hydrogen fuel cell vehicles, International Journal of Hydrogen Energy, https://doi.org/10.1016/j.ijhydene.2019.09.206

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Introduction Prior work demonstrated that current refueling costs contribute a large fraction of the total levelized cost of fuel for light-duty hydrogen fuel-cell electric vehicles (HFCEVs) [1,2]. Current small-capacity (~200-kg/day) refueling station costs are smallest when gaseous H2 is supplied to the refueling station in tube-trailers from a regional H2 producer (rather than in liquid form): The refueling station costs, amortized over the fuel pumped into vehicles over the station lifetime, are ~$6e7 per kg H2 in 2014 dollars. Hydrogen production from natural gas costs between $2 and $3 per kg H2, while transportation costs (which include large-scale storage, loading terminals, and trucking to the station) vary by transportation distance and delivered payload. For a carbon fiber overwrapped tube-trailer’s payload of ~1000 kg and a transportation distance of 100e300 mi, the transportation cost adds another $4e5 per kg H2. Thus, refueling-station costs contribute ~40e50% of the total fuel cost to the customer, currently at $12e15/kg H2. The refueling station's ~$6/kg-H2 cost contribution is apropos to early-deployment scenarios in which small singledispenser refueling stations deliver 200 kg H2/day using equipment manufactured with current technologies in small numbers. It is estimated that the station cost contribution can be reduced to approximately $2/kg H2 if the station capacity increases to 1000 kg/day (3e4 dispensers) while maintaining 80% utilization, if the station equipment is manufactured at high volume, and if cost reduction via learning by doing is assumed. Also, in a future scenario, where hydrogen is transported at scale to various markets via pipelines, the cost of transportation can be reduced to ~$2/kg. Considering the $2/kg-H2 production cost via steam methane reforming, both the (future) $2/kg-H2 refueling cost and the ~$2/kg pipeline transportation cost will need to be reduced to achieve the U.S. Department of Energy’s (DOE’s) long-term cost goal of $4/kg H2 at the dispenser. The refueling station costs are currently dominated by compressor, storage, and refrigeration costs, which account for approximately 50%, 20% and 12% of total equipment cost, respectively, in both the near-term 200-kg/day scenario and the future 1000-kg/day scenario. A refueling station schematic, Fig. A.1, showing major components is presented in an appendix. The costs of these three components alone contribute approximately $5/kg for the near term and $1.6/kg for the future scenarios [1]. The scenarios just described utilize high-pressure (700-bar) tanks on board the vehicle. The high-pressure Type IV tank lining is limited to 85  C, which can be exceeded if the H2 is not cooled to 40  C prior to dispensing. Note that the JouleThomson coefficient for H2 is negative for temperatures and pressures encountered during refueling, so the H2 temperature increases as the H2 expands into the vehicle’s tank from the high-pressure supply: Cooling costs are significant in the 700-bar tank scenario. This report explores potential refueling-station cost reduction if lower-pressure onboard storage options that operate near ambient temperature can be realized. In particular, refueling scenarios are considered for onboard hydrogen

storage in tanks containing metal hydride (MH) beds, assuming that these tanks can be developed and can meet specific performance and cost targets [3]. It will be shown that refueling-station costs are likely to be reduced when onboard MH tanks are used instead of 700-bar tanks and, further, refueling-station considerations actually constrain some elements of the MH tank design and MH material property optimization. There are numerous papers describing MH and MH storage tanks, e.g. Refs. [4e9], and there are numerous papers that consider the cost of HFCEV operation, e.g. Ref. [10], but we are not aware of detailed analyses of implications for refueling station costs or detailed comparisons between refuelingstation costs for high-pressure and MH onboard tank scenarios. The major contributions of this paper are to describe such an analysis and to demonstrate that onboard MH scenarios most likely offer cost savings at the refueling station. Further, we identify an important area where additional engineering is required, namely, the dispenser-to-vehicle interconnect, and show that the refueling cost is sensitive to this item through sensitivity to labor costs. Low-enthalpy MH materials are preferred for onboard storage because they may avoid parasitic consumption of H2 to release the H2 from the MH. Our work demonstrates that such a scenario is consistent with a refueling station that cools against ambient temperatures.

Methods e models and computations Metal hydrides Metal hydrides are useful for hydrogen storage because of their reversible chemical reactions with H2. The generic reaction [11] is M þ x/2 H2 % MHx þ Q where M is a metal or alloy and Q is the heat of reaction. Charging is exothermic, but discharging is endothermic. The equilibrium hydrogen gas pressure, PH, depends upon temperature, T, and the hydrogen concentration in the solid, C [12]. Isotherms of PH vs. C rise rapidly to a nearly constant plateau over a broad range of C. Beyond the plateau, PH rises rapidly again. The plateau region enables charging and discharging at constant PH. The plateau value of PH, designated Peq, depends upon T and the broad behavior is given approximately by the van’t Hoff equation, 
ln Peq P0 ¼  DS = R þ DH = RT;

(1)

where DH and DS are the standard enthalpy and entropy changes in the reaction, P0 is 1 bar, and R is the universal gas constant [12]. The bed charges when the headspace pressure exceeds Peq and discharges when less than Peq. Thus, the tank is controlled via its pressure and temperature. The rate of charging and discharging is determined by the T- and P-dependent chemical kinetics. Since Peq increases with T, the pressure drive for the absorption, P-Peq, decreases with temperature. This decrease counters kinetics factors that increase with temperature. As a result, the overall kinetics displays a peak; hence, the MH bed temperature must be kept within a limited temperature band of approximately 30 Ke40 K or the recharging rate will be impaired [13].

Please cite this article as: Frank ED et al., Refueling-station costs for metal hydride storage tanks on board hydrogen fuel cell vehicles, International Journal of Hydrogen Energy, https://doi.org/10.1016/j.ijhydene.2019.09.206

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international journal of hydrogen energy xxx (xxxx) xxx

10.0 0.5" ID

1000.0

8.0

1.25" ID

7.0 6.0

1.75" ID

100.0

5.0

2.0" ID Heat rate, T=30K

4.0

Heat rate, T=40K

10.0

3.0 2.0

1.0

1.0 0.1

0.0

0

5

10

15

20

25

30

35

Coolant mass rate (50 wt% Ethylene Glycol), kg/s

Fig. 1 e Pressure drop summed across 3-m-long coolant supply and 3-m return lines for several diameters of rubber line for 50 wt% EG at 60  C. The heat rate for the various flows is also plotted for 30 K and 40 K coolant temperature rises.

line plus 3-m return line was computed via the D’Arcy-Weisbach equation for rubber tubing with absolute roughness of 0.046. The Swamee-Jain equation was used to estimate the friction factor. The Reynolds number varied between 30,000 and 1,500,000 for 1.5 kg/s to 18 kg/s flows in tubing ranging from 0.500 to 2.000 ID. The corresponding friction factors varied from 0.020 to 0.029. Tables 1 and 2 show the fluid properties and the coolingfluid rates that are required in two refueling scenarios (presented below), one requiring 0.6 MW and the other 1.1 MW. If the pressure drop is limited to 20 psi total, summed over the supply and return lines, then lines with at least a 1.2500 ID are needed for the 0.6-MW scenario. For the 1.1-MW scenario, the line would need to have a 1.7500 to 2.0” ID to stay below 20 psi. Practical considerations for the operator will limit the weight and stiffness of the refueling interconnect. The refueling interconnect will require a cold-side coolant supply line, a hot-side coolant return line, plus the H2 supply line. For

Metal hydride refueling station assumptions: cooling interconnect and station attendant

10.0 0.5" ID

9.0

0.75" ID

1000.0

1.0" ID

8.0

1.25" ID

7.0

1.5" ID

6.0

1.75" ID

100.0

5.0

2.0" ID Heat rate, T=30K

4.0

Heat rate, T=40K

10.0

Heat Rate, MW

Feed + Return Pressure Drop, psi

10000.0

In the onboard refueling scenarios considered here, cooling is accomplished via fluid circulated between the vehicle and a heat exchanger (HX). The enthalpy range of 27e41 kJ/mol H2 corresponds to 0.38e0.57 MW average cooling duty when 5 kg of H2 is stored in 3 min (DOE targets). Since the temperature change in the fluid is limited to approximately 30 K, substantial heat capacity rates (Cp*dm/dt) will be required to achieve the high cooling duty. Although some authors express concern about cooling with water, we consider 50 wt % ethylene glycol (EG) because of its relatively high heat capacity and, at higher temperatures, we consider Therminol66 [16]. Figs. 1 and 2 examine the pressure drop in smooth-walled rubber lines of various diameters. Since three meters is approximately the length of the dispenser-to-vehicle dispensing lines, the pressure drop across a 3-m-long supply

1.0" ID 1.5" ID

Metal hydride tanks Metal hydride tanks are designed with careful consideration of thermal management. The rapid exothermic heat release during refueling translates into a high cooling duty of approximately 0.5e1 MW. Therefore, tank cooling is unlikely to be done on board via heat rejection to ambient air; rather, cooling via a liquid loop through the tank during refueling is assumed. The thermal management must be designed carefully because it must maintain a relatively narrow temperature range; e.g., one study required a sodium alanate bed to be maintained within a 30 K window [13]. Raju and Kumar [15] simulated heat transfer in sodium alanate beds. When cooling was mismatched to the heat generation rate in a 2D finiteelement multi-physics simulation, regions of the bed would be either too cool or too hot to support high kinetics, thus reducing the total hydrogen charge that could be achieved within a target fueling time. Across a number of sensitivity variables, the best performance in Ref. [15] involved a cooling fluid temperature at approximately 30  C below the final bed charging temperature. Given these considerations, one cannot use a large temperature difference between bed and coolant inlet to increase the capacity for carrying away heat. The alternative is to increase the cooling fluid flow rate, which we now examine.

9.0

0.75" ID

Heat Rate, MW

10000.0 Feed + Return Pressure Drop, psi

The range of possible enthalpies is not very large for materials that meet DOE targets. The DOE target of 5-bar discharge pressure between 40  C and 80  C, the relatively small range of values for the entropy change, and the van’t Hoff equation jointly imply that the hydriding enthalpy change is in the range of 27e41 kJ/mol H2 [14]. When MHs require discharge temperatures higher than the waste heat temperature of the vehicle fuel cell, hydrogen must be used for heat to desorb the H2, thus reducing vehicle range and whole-pathway energy efficiency. Thus, highenthalpy MHs are disfavored. Rather than choosing an arbitrary upper cutoff, we explore the full range of enthalpies in this study, as explained below.

3.0 2.0

1.0

1.0 0.0

0.1

0

5

10

15

20

25

30

35

Coolant mass rate (Therminol-66), kg/s

Fig. 2 e Pressure drop summed across 3-m-long coolant supply and 3-m return lines for several diameters of rubber line for Therminol-66. The heat rate for the various flows is also plotted for 30 K and 40 K coolant temperature rises.

Please cite this article as: Frank ED et al., Refueling-station costs for metal hydride storage tanks on board hydrogen fuel cell vehicles, International Journal of Hydrogen Energy, https://doi.org/10.1016/j.ijhydene.2019.09.206

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Table 1 e Properties of cooling fluids in this study. To be conservative in hand computations, Cp and density values are the lowest values in the coolant temperature ranges while the viscosity is the highest. The temperature ranges were 60e100  C (EG) and 130e168  C (Therminol-66). Aspen Plus simulations used full, temperature-dependent material properties. EG Ethylene Glycol.

Cp Density Dynamic viscosity

J/kg.K kg/m3 mPa.s

50 wt% EG

Therminol-66

2825 1023 1.55

1964 908 1.9

60  C case for this study because the exit temperature for the 40  C case may be too cold for efficient cooling against ambient temperatures during the summer. The heat rate varied over the course of the simulated refueling. The maximum, minimum, and average heat rates over the refueling cycle are plotted in Fig. 4. The maximum heat duty for this scenario was 0.5 MW. Another 0.1 MW was added for heat associated with H2 dispensing (enthalpy of H2 entering the tank). Summary of low-enthalpy, low-temperature scenario: 0.6 MW, Tin ¼ 60  C, Tout ¼ 100  C, P ¼ 10 MPa, Tair ¼ 38  C.

Scenario 2: high enthalpy, high temperature comparison, a common gasoline dispenser utilizes a single line with ¾” ID and 1-1/800 OD. Additional coolant line thickness will be required for insulation, which will increase the OD. The H2 refueling interconnect seems to require two lines, each with at least approximately 1.2500 to 2.0” ID, plus the H2 line itself. This brief coolant supply analysis leads to several conclusions. First, a conservative refueling station cost estimate should consider wages for an attendant: Self-service may not be possible because of multiple bulky hoses. Second, coolant supply requires further detailed analysis, e.g., the 20-psi limit provides an illustrative example, but is unjustified otherwise.

Scenarios We examined two scenarios that are illustrative of the key considerations and which bracket the range of DH and DT.

Scenario 1: low enthalpy, low temperature One study [17] reverse-engineered the material properties that are needed to achieve DOE targets. The study required that no H2 be used for heat during H2 desorption. The waste heat temperature of an 80-kWe low-temperature fuel cell thereby constrains T and DH when the tank must maintain at least 5 bar during discharge. The study concluded that a MH with DS ¼ 110 J/mol.K, DH ¼ 28 kJ/mol, and activation energy Ek ¼ 45 kJ/mol was needed to meet the DOE targets. The charging pressure was 100 bar and the optimum bed temperature was equal to 97  C. The coolant inlet temperature, mass flow rate, and peak cooling duty must be determined to size the HX for the present analysis. The coolant-tank exit temperature was examined as a function of time for 40 , 60 , and 80  C inlet temperatures and for several coolant flow rates. Fig. 3 shows simulation results when the inlet temperature was 60  C. We used the

Johnson et al. [18] performed experiments with a full-scale tank utilizing sodium alanate. The tank had a usable capacity of 3 kg H2, achieved 2 g/s desorption, and satisfied the Federal Test Protocol-75 (FTP-75) and GM Rosso drive cycles. The project did not focus on DOE density targets or recharging time targets, which were not achievable with sodium alanate. Instead, the project focused on scale-up issues, like accuracy of modeling, mass transfer, and heat management in a fullscale system. The experiments achieved 3.2 wt% H2 after 10 min and 4.0 wt% H2 after 30 min of charging. Although higher-temperature MHs are disfavored because of onboard H2 use for desorption, this system is interesting for a refueling-station scenario for two reasons. The sodium alanate system has two reactions with enthalpies of 37 kJ/mol H2 (tet reaction) and 47 kJ/mol H2 (hex reaction) or a weightaveraged enthalpy of 40 kJ/mol H2 [13]. Thus, this system lies at the upper end of the 27e41 kJ/mol-H2 enthalpy range, which complements the low-enthalpy scenario, bracketing the likely range of enthalpies. Also, the two reactions occur at different temperatures and with different heat rates, introducing a complication not faced in the simpler low-enthalpy model. We constructed a refueling-station scenario from this information as follows: Figure 14 in Ref. [18] shows the state of charge, pressure, and bed temperature during a 20-min charging cycle. Most (85%) of the charging occurs during the first 10 min. This portion of each curve was scaled to 3 min and a total charge of 5 kg H2. Based upon this scaling of the

Table 2 e Mass rates for various cooling fluids. EG Ethylene Glycol. Heat Duty, MW 0.6 1.1

DT,  C

50 wt% EG, kg/s

Therminol66, kg/s

30 40 30 40

7.1 5.3 13 9.7

10. 7.6 19 14

Fig. 3 e Coolant temperature at the MH tank outlet vs. time for various coolant flow rates in the model of Ref. [17] when the inlet temperature is 60  C. Tf out ¼ coolant temperature at the tank outlet; Tf in ¼ coolant temperature at the tank _ f ¼ coolant mass flow rate. inlet; m

Please cite this article as: Frank ED et al., Refueling-station costs for metal hydride storage tanks on board hydrogen fuel cell vehicles, International Journal of Hydrogen Energy, https://doi.org/10.1016/j.ijhydene.2019.09.206

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Heat exchanger design and cost estimation

Fig. 4 e Peak (solid), average (dotted), and minimum (dashed) heat duties for various coolant inlet temperature scenarios considered as a function of coolant mass flow rate. The peak duty is substantially larger than the average duty (0.36 MW) and was used to size the HX.

Johnson et al. temperature graph, the hex phase charging occurs for 46 seconds at 210  C bed temperature, accounts for 38% of the stored H2, and has a heat duty of 0.97 MW. The tet phase charging then lasts 134 seconds at a bed temperature between 180 and 165  C and has a heat duty of 0.43 MW. An additional 0.1-MW heat duty is added for the H2 dispensing, giving a total heat duty of 1.1 MW (hex) and 0.53 MW (tet). The H2 supply pressure was 13.8 MPa. With regard to coolant flow, Johnson et al. [18] used Xceltherm-600 heat exchange fluid fed at 130  C to the tank inlet. Pressure drop across the tank limited flow to 8.2 L/s. Supplying Xceltherm-600 at 8.2 L/s works for a 20-min charging cycle, but for a 3-min recharging time, Xceltherm600 with heat capacity of 2.12 kJ/kg.K at 8.2 L/s would exit with T ¼ 206 K, which is too high to extract heat from a bed that is between 200 K and 220 K. For the present analysis, we assume that the tank can be redesigned to tolerate the higher coolant pressures associated with higher fluid flow rates. A material model for Xceltherm-600 was not available for use in Aspen Plus, so we used an alternative cooling fluid, Therminol-66, in our work. In our Aspen model, a 14-kg/s flow was required. Johnson et al. [18] studied charging when the tank was partially discharged and also when the tank was cold, and determined that the 130  C cooling fluid could heat the cold tank enough to enable the hydriding reaction to begin. The tank warmed to the ideal recharging conditions quickly enough via reaction heat that the total recharging time increased only slightly. Thus, we neglect external, supplementary heating since its use would be limited and the costs would likely be small vs. the air-cooled heat exchanger (ACHX). Future design work can revisit this assumption.

Summary of high-enthalpy, high-temperature scenario (Tair ¼ 38  C, P ¼ 13.8 MPa)
Hex: 1.1 MW, Therminol-66, Tin ¼ 130  C, Tout ¼ 168  C
Tet: 0.53 MW, Therminol-66, Tin ¼ 130  C, Tout ¼ 148  C

Since water use is a concern for renewable-energy technologies, ACHXs were considered for cooling during refueling. Consideration of ASHRAE climate tables (0.4% dry bulb cooling temperature) showed that the high- and low-temperature scenarios were hot enough for ACHX cooling against ambient conditions throughout the year in most U.S. locations. ACHX sizing and cost estimation were approached in several ways for the sake of verification, namely Aspen Plus, the study by Loh et al. [19], and vendor quotes. Vendor quotes give the most up-to-date information, but do not give insight into sensitivity. Aspen Plus provides both sizing and cost information and can be used to study the sensitivity of the sizing estimate. Ref. [19] does not provide sizing estimates, but can be used to get an alternative cost estimate from the ACHX area computed in Aspen Plus. ACHXs were specified and studied in Aspen Plus. The ACHX inlet and outlet temperatures were fixed to match the recharging conditions defined above for the high- and lowenthalpy scenarios. The mass flow rate at the inlet was fixed by the Cp and temperature swing of the fluid to meet the target heat duty. Aspen Plus allowed many variables to be specified, such as tube geometry, number of rows, number of passes, fin type, fin size and spacing, and air mass flow rate. Several configurations were examined and were assessed for their ability to meet the cooling duties in the two scenarios. Since this is an initial assessment of the potential for MH to reduce refueling-station costs, and since ACHX cost is relatively insensitive to size variation at this scale, the design space was not examined exhaustively, nor was formal optimization attempted, but ACHX area was examined in a sensitivity analysis (see Results, Results). Aspen Plus provided capital and operating cost estimates. For comparison, the total bare tube area computed by Aspen Plus was used to compute the ACHX cost from Ref. [19]. The result from Ref. [19] is in 2008 dollars, FOB at the manufacturer gate, which was adjusted to 2018 dollars via the Chemical Engineering Plant Cost Index (CEPCI). The results given by Hudson [20] were used to obtain a second estimate for the fan power for the Aspen Plus designs. After Aspen Plus was used to explore the physical aspects of the ACHX (discussed below), several vendors were asked for budget-level cost estimates, and two of them responded. The vendors’ cost estimates are presented in Results.

Techno-economic analysis The techno-economic analysis (TEA) was performed using the Hydrogen Delivery Scenario Analysis Model (HDSAM), V3.0 [21]. HDSAM has been developed over the past 15 years with support from the DOE. It uses an engineering bottom-up economics approach to calculate the cost of H2 delivery. Details are in Ref. [22], but, in brief, HDSAM computes the fuel selling-price contributions required to achieve a zero net present value for a cash flow model for a target internal rate of return (IRR). The HDSAM financial assumptions (Table 3) were chosen to be parsimonious with other DOE transportation analyses to facilitate comparisons. A description of the

Please cite this article as: Frank ED et al., Refueling-station costs for metal hydride storage tanks on board hydrogen fuel cell vehicles, International Journal of Hydrogen Energy, https://doi.org/10.1016/j.ijhydene.2019.09.206

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Table 3 e Summary of TEA financial assumptions. Parameter Reference year dollars Nominal IRR Real after-tax discount rate Analysis period, yr State taxes Federal taxes Inflation rate Equity ratio

Value 2016 12% 10% 30 6% 21% 1.9% 100%

HDSAM tool and its documentation are publicly available for download from the tool’s web site, https://hdsam.es.anl.gov/. HDSAM considers all costs from the gate of a centralized production facility to delivery at a station and dispensing into a vehicle, including distribution by truck or pipeline and any required storage and conditioning, but costs associated with H2 production are outside the scope of the model. HDSAM provides city population and fleet data and is supplied with assumptions such as market type (urban, rural, interstate, or combination), population and vehicle ownership, hydrogen vehicle penetration rate, H2 transmission mode (gaseous tubetrailer, liquid H2 truck, or gaseous pipeline), terminal storage modes (geological or liquid), refueling station capacity, and dispensing mode (350 bar, 700 bar, etc.). Given these assumptions, HDSAM computes the number and types of various items of equipment such as compressors, liquefiers, dispensers, and storage equipment and optimizes the choices with respect to cost. Examples of optimization include optimization of compression and storage equipment [2,23]. HDSAM uses scaling rules for equipment costs as a function of size. Scaling rules are also available to estimate future cost reductions that are likely to accrue with increased production experience. Key industrial collaborators participated in HDSAM review and development, e.g., provided various rules of thumb, gave access to their public data, reviewed project documents, and participated in beta-testing of the model. The capital cost of each fueling component was acquired from one or more vendors by providing detailed engineering and performance specifications for each component. The costs of components that are common with other onboard storage fueling, e.g., storage tanks, compressors, and dispensers, have been documented in previous publications [7,22]. The costs of additional components that are unique to fueling MH tanks, e.g., HXs, fans, and pumps, are described in detail below. We evaluated the case of 4% H2 HFCEV market penetration for an urban setting equal in population to Sacramento, CA, and assumed vehicles travel 12,000 mi/yr with average fuel economy of 55 mi/gallon-gasoline-equivalent. We assumed that summer demand surges by 10% for 120 days. These factors help determine the size of the regional terminal and infrastructure bulk storage. Pipeline distribution is not economical with a market of this scale; therefore, tube-trailer trucks were assumed for distribution. While many station models are possible in HDSAM, the present analysis examines a 1000-kg/day station, near the upper end of station size consistent with tube-trailer distribution. The station utilization was assumed to be 80%, hose occupancy fraction was 50% during peak demand hour, fill

time was 3 min, linger time was 2 min, and 8% surge in demand was assumed on Fridays. This set of assumptions models a near-to mid-term scenario that was selected because it is not too far in the future, yet is not dominated by the poor economics associated with a single dispenser and low utilization. Under these assumptions, the station required three dispensers and dispensed approximately 300,000 kg/yr. Equipment sizing is critical because capital expenses dominate the result: Mitigating oversizing for peak demand is important. Geological storage near the production site mitigates demand variation at the terminal, while fast-peaking at the refueling station is mitigated via an on-site high-pressure cascade storage buffer. The cascade storage buffer between the tube-trailer and the vehicle reduces the compressor capacity requirement, and thus the refueling costs [2,23]. Some major refueling components, e.g., dispensers and H2 compressors, are evolving technologies for which cost reductions are expected because of learning and scale-up. We applied mid-level savings factors to current-day cost estimates for these items: Costs for items with significant, moderate, and limited industrial production experience were reduced by 79%, 61%, and 47%, respectively [22]. The ACHX and coolant pump costs that are required for the MH scenario, however, were not reduced, partly to be conservative, but also because these are fully mature technologies that are in widespread use.

Results An ACHX design for the low-temperature, low-enthalpy scenario was first explored using propylene glycol (PG). The design was challenging because of the moderate temperature difference between PG and ambient and because of the high viscosity of PG. Initial designs led to laminar flow, low thermal conductance (U) factors, and high area. Reducing the tube count and reducing the tube ID increased the fluid velocity, which improved the hot-side film coefficient, and configuring the ACHX with six rows in six passes in cross-counterflow improved the mean temperature difference. Nevertheless, a substantial area was required (848 ft2 bare tube). The PG Reynolds numbers indicated turbulent flow at the inlet (5300) and laminar flow at the outlet (1350). The effective bare-tube area U-factor was 43 BTU/h.ft2. F and the total-area U-factor was 1.9 BTU/h.ft2. F. PG was originally favored because of its high heat capacity, but EG has fourfold lower viscosity at the HX outlet temperature. Since lower viscosity may improve the hot-side film coefficient through increased turbulence, we examined EG. Employing EG reduced the ACHX area twofold, increased the bare-area Ufactor to 118 BTU/h.ft2. F, and increased the total-area U-factor to 5.3 BTU/h.ft2. F. The rest of the discussion therefore focuses on EG as the cooling fluid for the low-temperature case. Given these findings, initial budget estimates were requested from vendors for coolers at 38  C ambient temperature using EG for the low-temperature case and Therminol66 for the high-temperature case (to avoid phase changes without going to elevated pressure). The vendors were also asked to consider daily and seasonal temperature variations similar to those for Washington, D.C. Vendor #1 provided the

Please cite this article as: Frank ED et al., Refueling-station costs for metal hydride storage tanks on board hydrogen fuel cell vehicles, International Journal of Hydrogen Energy, https://doi.org/10.1016/j.ijhydene.2019.09.206

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ACHX cost including $5k for a variable-frequency drive (VFD), temperature transmitters, and other required items for handling temperature variations. Vendor #2 did not include these items, but provided the cost of an ACHX with a VFDcompatible motor. We added $5k to the Vendor #2 estimate to compare with the Vendor #1 estimate. Vendor #2 also offered walkway and hail-guard options, which added $3.7k and $0.59k, respectively. We included these options in the Vendor #2 cost to be conservative in our analysis. Table 4 summarizes the costs from Aspen, Ref. [19], and the two ACHX vendors. The cost from Ref. [19] was computed via the Aspen Plus bare area and then adjusted for inflation between 1998 and 2018 via the CEPCI as a crosscheck on the Aspen pricing. The Aspen Plus and Ref. [19] estimates have $5000 in 2018 dollars added to them for temperatureregulation equipment. Several fan power estimates were available. Vendor #1 provided only the nameplate ratings of its fan motors. Vendor #2 provided the pressure drop across the ACHX, the airflow, and the vendor’s estimated fan power after accounting for gearbox/belt efficiency and fan efficiency, but not motor efficiency. The vendor did not explain how the power was estimated, so we computed the fan power for the vendor’s stated airflow and static pressure change assuming 65% fan efficiency, 95% belt/gearbox efficiency, and 90% motor efficiency. Hudson [20] provided an independent estimate of the minimum and maximum fan power required for an ACHX, based upon the bare area and the number of rows. Given these data (Table 5), the Hudson powers were selected as the most defensible. They are comparable to or slightly above the pressure-drop-based powers and are also empirical values suggested for initial design by Hudson. We used the maximum Hudson fan power, after rounding to one digit, for the TEA analysis, i.e., 7 HP total electrical power for the lowtemperature case and 6 HP for the high-temperature case. Fan power will be examined in the sensitivity analysis.

The cooling-fluid circulation power is another operating cost that must be considered. It is difficult to estimate because the dispenser coolant line diameters and the H2 tank pressure drops are not known and are a matter of system optimization. We assumed 20 psi across the coolant supply and return lines. Both vendors estimated a 5-psi drop across the ACHX for the low-enthalpy case. For the high-enthalpy case, one vendor estimated 7 psi across the ACHX and the other estimated 10 psi. We adopted a 10-psi drop across the ACHX in the highenthalpy case for operating cost calculations and, lacking a station site layout, assume that the pressure drop in the station piping between the dispenser and the ACHX is designed to a budget of 5 psi. The pressure drop across the onboard tank must now be estimated. Johnson et al. [18] reported a 6-psi drop across the full-scale sodium alanate tank with Xceltherm-600 heat exchange fluid and a flow of 1.6 kg/s in each of the four (parallel) modules, for a total flow of 6.4 kg/s. The heat duty in the time-scaled scenario considered here requires 15 kg/s. If the pressure drop scales with the square of the velocity, the pressure drop across the tank in the scaled system will be 33 psi. Johnson et al. [18] commented that half of the pressure drop occurred in an unoptimized fluid manifold. Assuming a 50% improvement in the drop across the manifold, we assume that the pressure drop across the tank in the high-enthalpy case can be limited to 25 psi. No information is available for the pressure drop across the tank in the low-enthalpy case, but the 4.3-kg/s flow in the lowenthalpy case would reduce the pressure drop seen in Ref. [18] to 1.8 psi because of the 67% velocity reduction and 1.23-fold density increase. The change in Reynolds numbers between the Xceltherm-600 and EG flows implies different friction factors, but we found this effect to be small and consider only the effect of velocity on the pressure drop. Table 6 displays the various contributions to the coolant circulation pressure drop and the resulting power demand.

Table 4 e ACHX design and costs. NP¼Not provided. Source

Coolant dm/dt, kg/sa

Total HX area, ft2

Low-enthalpy, low-temperature Aspen 4.3 Ref. [19] Vendor 1 4.3 Vendor 2 4.3 High-enthalpy, high-temperature Aspen 14 Ref. [19] Vendor 1 15 Vendor 2 15 a b c

d e

8300 3800 7300 9100 3800 4900

Bare tube area, ft2

Tubes/ rows/ passes

U, total,b BTU/ h.ft2. F

U, bare, BTU/ h.ft2. F

Footprint, LxW, ft

Cost, k$c

370 370d 290e 310

156/6/6 NP 224/4/NP 130/4/6

5.3

118

14  6

10.4 5.4

129

88 12  6

29 51 62 43

550 550d 290e 310

120/4/4

2.4

40

16  7

224/4/NP 158/4/2

6.4 5.7

91

88 12  5

33 57 60 42

Cooling fluid supply rate. U factor based upon total finned area. 2018 dollars. ACHX FOB at the gate of the manufacturer. Does not include field costs, e.g., installation, support structure, and ladders. Ref. [19] results were adjusted from 1998 to 2018 by a factor of 1.56 via the CPI, https://www.bls.gov/data/inflation_calculator.htm/. Assumed area, taken from Aspen. Not provided by vendor. Estimated from vendor’s drawing (tube length) and vendor’s specifications for tube diameter and tube count.

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Table 5 e ACHX properties, air-side. The air inlet temperature (ambient) is 38  C. Tout ¼ Air temperature at the outlet; Vface ¼ Velocity at the inlet; DP ¼ Static pressure drop; DP Power ¼ Power computed from the pressure drop, including inefficiencies; Vendor power ¼ Power computed by the vendor, including inefficiencies; Hudson ¼ power computed from the bare area and row count via Ref. [20]; Vendor nameplate ¼ Sum of motor nameplate power ratings specified by vendor; Blank cells ¼ Information not provided by the source. Source

Airflow, ACFM

Tout,  F

Vface, m/s

DP, Pa

DP Power, HP

2.6

67.3

2.6

3.1

123

6.3

3.7

102

6.9

2.5

121

4.5

Low-enthalpy, low-temperature Aspen 34041 159 Vendor 1 43037a 145 Vendor 2 45472 145 High-enthalpy, high-temperature Aspen 59634 162 Vendor 1 42886a 184 Vendor 2 32668 214 a b

Vendor power, HP

Hudson min, HP

Hudson max, HP

Vendor nameplate, HP

9.8b

4.7 4.3 5.0

5.8 5.8 6.8

15 10

5.8b

8.1 4.3 4.6

11 5.8 6.2

15 10

SCFM. Increased to reflect 90% motor efficiency.

value, 8 h/day, is equal to running only when the H2 is being dispensed. The ACHX cost variation was computed by considering ACHX cost as a function of bare area. Cost scaling factors were estimated via the cost curve in Ref. [19] as follows: The average bare area from the vendors was 300 ft2. The costs for half-area (150-ft2), baseline (300-ft2), and doubled-area (600ft2) ACHXs were obtained from Ref. [19] and the ratios of Ref. [19] costs for half-area and doubled-area to the baseline area were obtained, namely, 0.91 and 1.2, respectively. The vendor ACHX cost in Table 7 was multiplied by these scaling factors to obtain half-area and doubled-area scenario costs, namely, $46,400 and $61,200 per ACHX. These ACHX costsensitivity scenarios were run with the baseline fan and pump powers so that the effect of ACHX cost could be seen separately. Finally, the sensitivity to the dispenser cost was examined because the two MH scenarios required very different coolant rates and because the MH dispenser will be more complicated than the 700-bar compressed-H2 dispenser. Unfortunately, we have no data to estimate the cost differences between these options. We therefore mapped out the refueling cost for dispenser cost multipliers between 0.5 and 2.0 in steps of 0.25, found that the refueling cost changed linearly, and report values for half and doubled cost for the dispensers (Table 8).

Given the uncertainties in estimating the pressure drop, coolant pump power will be examined in the sensitivity analysis. The results above determined the key inputs to the TEA, as summarized in Table 7. The ACHX cost is the uninstalled cost of the ACHX, FOB at the manufacturer, averaged over the two vendors. The ACHX installation cost factor was equal to two. Pump costs in 1998 dollars were obtained from Ref. [19] for 100-gpm and 300-gpm centrifugal pumps and were corrected to 2018 via the CEPCI. The coolant circulation pump operation time is equal to the refueling time. The ACHX fans were assumed to operate 100% of the time during the day and 50% during the night, totaling 18 h/day. The sensitivity analysis will examine the fan operation time. Fig. 5 shows the contributions to the total delivery cost per kg of H2 for a baseline scenario (700-bar pressurized tank) and for the two MH scenarios. Delivery costs do not include H2 production. The costs associated with delivering H2 to the refueling station, including the cost of the gaseous H2 (GH2) terminal, were kept the same for all cases. Fig. 6 breaks down the refueling station costs from Fig. 5. Sensitivity analysis examined the ACHX cost, ACHX fan power, coolant pump power, ACHX fan running time, dispenser cost, labor cost, and dispensing pressure. Since the high-enthalpy scenario had a slightly higher refueling cost, the high-temperature case was selected as the baseline for the sensitivity analysis. Equal-probability confidence intervals are not known for the various parameters, so they were varied by a factor of 2 in most cases to explore the associated rate of change in the refueling cost. The ACHX run-time minimum

Discussion The high-temperature scenario was challenged by the presence of two MH phases with different charging rates,

Table 6 e Estimated pressure drops in the cooling-fluid circulation loop. Scenario Low-enthalpy High-enthalpy a b c

Dispenser lines,a psi

Tank, psi

Station piping, psi

ACHX, psi

Total, psi

dm/dt,b kg/s

Power,c HP

20 20

1.8 25

5 5

5 10

32 60

4.3 15

1.8 10

Sum over coolant supply and return lines between the vehicle and the dispenser. Fluid mass flow rate. Includes pump efficiency (75%) and motor efficiency (90%).

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Table 7 e Summary of TEA inputs. Scenario Low-Enthalpy High-Enthalpy a

H2 fill pressure, bar

ACHX cost, 2018 $

Pump cost, 2018 $a

ACHX Fan Power, kW

Circulation Power, kW

100 138

52,000 51,000

6500 7700

5.1 4.4

1.3 7.4

Adjusted to 2018 from 1998 via the CEPCI.

Fig. 5 e Contributions to the total delivery cost per kg of H2 for a 700-bar pressurized tank and for MH tanks utilizing either high-temperature, high-enthalpy MH or low-temperature, low-enthalpy MH. All values are 2016 dollars.

Fig. 6 e Contributions to the refueling station cost per kg of H2 for a 700-bar pressurized tank and for MH tanks utilizing either high-temperature, high-enthalpy MH or low-temperature, low-enthalpy MH. Electrical Installation includes only the costs for electrical installation and on-site wiring. Electrical power operating costs in the figure are lumped to the components that consume power, namely, Compressor and Cooling. All values are 2016 dollars.

enthalpies, and bed temperatures. As a result, the average heat rate and peak heat rate were quite different, by approximately a factor of 2, which led to challenges in the HX design. In our Aspen studies, the high-temperature, high-enthalpy case was approached by first designing an ACHX for the early hex phase charging when the heat duty was highest. We then examined how the ACHX performed during the lowertemperature, lower-heat-duty tet phase charging. When the ACHX designed for the hex phase was evaluated under the

lower temperatures appropriate for the tet phase charging, the ACHX overcooled the fluid: The return temperature (target 130  C) was 116  C. The 130  C target could be achieved by reducing the air mass flow rate by a factor of 3. High-enthalpy systems operate at higher temperatures, which led to smaller HXs than in the low-enthalpy scenario, but several challenges resulted from the higher temperature. The temperature was too high for PG stability, and 92 wt% EG required approximately 30 psi (gauge) to remain liquid. For

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Table 8 e Sensitivity analysis results. For attendants, the number of people is the average number of attendants allocated to the refueling operation at work simultaneously over the 18-hour day. Parameter ACHX area, ft2 ACHX fan power, kW Coolant pump power, kW ACHX fan runtime, hr/day Fill pressure, MPa Attendants, # of people Dispenser cost,a k$ (2015) a

Parameter Value

Refueling Cost, $/kg

Low

Baseline

High

Low

Baseline

High

150 2.2 3.7 8 10 0 69

300 4.4 7.4 18 13.8 0.46 140

600 8.8 14.7 24 20 1 280

$1.21 $1.21 $1.23 $1.21 $1.19 $1.08 $1.16

$1.23 $1.23 $1.23 $1.23 $1.23 $1.23 $1.23

$1.27 $1.26 $1.24 $1.24 $1.28 $1.40 $1.38

Dispenser cost is the total uninstalled cost of all three dispensers.

comparison, the sodium alanate tank in Ref. [18] was limited to 7 psi because of its particular shell design. This constraint can be avoided with alternative designs, but at the expense of weight and volume. We utilized Therminol-66 in our model to avoid pressure and phase change, but Therminol-66 has lower heat capacity and higher viscosity than the water-based coolants. Thus, the high-enthalpy systems require large cooling-fluid mass flow rates, which lead to larger pressure drops. If constraints preclude increasing the diameters of the various coolant lines, then it is unclear whether the Therminol-66 or the EG will have a higher net operating pressure despite the need to elevate the EG pressure to avoid phase changes. The low-enthalpy and high-enthalpy scenarios had similar cost results. The cost savings for H2 refueling between the 700-bar pressurized tank and the less expensive MH scenarios arose from reduced compression cost, $0.71/ kg H2 (low-enthalpy) or $0.68/kg H2 (high-enthalpy). The high-pressure storage cascade provided a small savings of $0.07/kg H2 (low-enthalpy) or $0.05/kg H2 (high-enthalpy). Note that bulk storage of H2 at the station is provided by the tube-trailers: The high-pressure storage cascade is a secondary component that only stores enough H 2 to smooth out hourly peaking and is used to reduce the compressor size. Its size was varied to optimize each scenario, but its effect is moderate. The cooling costs were approximately equal for cooling 700-bar H2 to 40  C prior to dispensing and for cooling the MH bed during recharging. In total, the MH scenarios reduced the $2.2M (2016 dollars) refueling station capital cost by $0.9M (low-enthalpy) or $0.8M (highenthalpy). Sensitivity analysis showed that the results were relatively insensitive to a twofold variation in most key parameters. The largest model errors relate to uncertainties in the cooling heat duty and in cooling-fluid pressure drops. The former tie to costs through the ACHX area, but halving or doubling the ACHX area only affected the dispensing cost by 2e4 cents. This insensitivity occurs because the station ACHX size is so small compared to industrial-scale coolers that the ACHX cost curve is dominated by a pedestal. Uncertainties in pressure drops had almost no effect on fuel cost because the coolant pump is a relatively minor cost and its associated power is small compared to compressor power. The largest contributors to uncertainty in costs were the filling pressure and the station attendant labor. The low-

enthalpy and high-enthalpy fill pressures (10 MPa and 13.8 MPa, respectively) require different compressor and storage cascade costs, which correspond to a $0.04 difference in refueling cost. If, for some reason, 20-MPa dispensing were required, it would increase the refueling cost by $0.05 above the high-enthalpy case. It is difficult to estimate labor needs because of the uncertainty in the dispenser-to-vehicle interconnect, which may require a full-service station under the most pessimistic assumptions. To work around this uncertainty, we considered a range of labor needs that spans the possible space for comparing MH scenarios with high-pressure scenarios. In particular, we considered equal labor needs for high-pressure and MH scenarios and also considered requiring a full-time person dedicated to dispensing fuel. With regard to labor, filling stations in the U.S. almost always comprise a convenience store plus refueling dispensers. The underlying HDSAM model assumes that labor is shared between the refueling operation and the convenience store (all other costs and revenues related to the convenience store are outside the analysis scope). The attendants generally stay in the store and refueling is self-service. All of our prior analyses have allocated 1/3 of the station labor to refueling, as a conservative estimate. This method allocates 0.46 people for 18 h/day to refueling for the station modeled here and is responsible for $0.15/kg H2 of the refueling cost. This treatment does not allow for the attendant to provide full-service pumping. If the refueling interconnect is incompatible with self-service and if the refueling labor allocation increases to one person to allow a person to be at the pumps for all 18 h of daily operation, then labor costs contribute $0.32/kg H2. With regard to dispenser cost, note that the coolant pump was accounted for separately, so one large component of the difference in dispenser cost between the scenarios was already figured into the refueling costs. The twofold variation shown in Table 8 is likely overly pessimistic, yet even with such a large variation, the effect of dispenser cost on refueling cost does not outweigh the savings from reducing the compressor cost. If the dispenser cost differed by 25% between the scenarios, then the refueling cost would only change by 4 cents. We conclude that the dispenser cost is a good target for cost reduction in all scenarios, but uncertainty in dispenser cost is unlikely to change the result that MH tanks reduce refueling costs relative to 700-bar compressed-gas tanks.

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The scope of this analysis was limited to the refueling station costs. We did not evaluate vehicle cost differences for the different tank options. Vehicle cost differences could enhance or cancel the cost reductions identified here and must be considered in a total-cost-of-ownership analysis before choosing a technology. In addition, current vehicles and stations utilize high-pressure tanks. If additional time is needed to develop MH-based vehicles, more high-pressure stations will be deployed in the meantime. One must evaluate whether these stations can be converted to serve MH vehicles, e.g., whether adequate space for the ACHX is present, and the cost implications of converting. The ACHX footprint of 12 ft  4 ft (Table 4) is comparable to a vehicle footprint, suggesting that the forecourt area is roughly commensurate with the ACHX footprint. Future work could examine whether the ACHX can be placed on the forecourt canopy, and could examine noise levels. Refueling stations with off-board cooling will mix fluids between vehicles. The volume of cooling fluid in the offboard loop (return line, HX and associated piping, supply line) is substantial compared to the onboard volume. Even if each vehicle connects to a dedicated cooling circuit, that circuit will contain fluid from previous vehicles. Mixing fluids between vehicles has liability and acceptance implications. Purification systems may be required to prevent faulty vehicles from affecting other vehicles. The design and costs of these elements were not considered. On the other hand, if separate cooling circuits are not required, economies of scale may be possible for the ACHX by using one larger HX rather than one per dispenser. Using Ref. [19] to scale the vendor quotes, if one ACHX cooler 3 times the size of the three coolers in the current model is employed, then the refueling cost drops by $0.10. This result is only suggestive because additional controls and equipment would be required, and these were not considered. This analysis, to be conservative, assumed 38  C ambient temperature, a temperature at or above the ASHRAE dry-bulb temperature corresponding to 0.4% annual cumulative frequency for many of even the hottest locations in the U.S. Many locations will experience lower temperatures and will be able to use smaller ACHXs.

Conclusions We estimated cost savings for hydrogen distribution and dispensing into HFCEVs that might occur when MH storage tanks are used instead of high-pressure (700-bar) gas storage tanks. We considered a low-temperature, low-enthalpy scenario and a high-temperature, high-enthalpy scenario in order to bracket the design space. The MH scenarios achieved $0.71e$0.75/kg-H2 savings (in 2016 dollars) by reducing compressor costs without incurring the cryogenics costs associated with cold-storage options like metal-organic frameworks and cryo-compression. The refueling cost savings for MH tank scenarios were robust against most uncertainties. The largest sensitivities were to the tank pressure, which is determined by the MH, and to the station labor allocated to refueling. The cost of a

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full-service attendant, if the refueling interconnect prevents self-service, was the single largest cost uncertainty. The highenthalpy case required larger coolant lines, which may make it harder to design a self-service station. The refueling-station design is likely to provide constraints on the MH tank R&D. Although the pressure drops in the coolant lines had little effect on refueling cost, the tank must withstand the required pressure, affecting cost and weight. The temperature sensitivity of the MH kinetics and the coolant heat capacity jointly determine the coolant mass flow rate. The mass flow rate, coolant line diameter, and tank piping geometry determine the fluid velocity. The fluid velocity, in the context of the coolant boiling point relative to the MH charging temperature, determines the pressure seen by the tank and affects the refueling interconnect physical design. Thus, the need to avoid selfservice has implications for the MH kinetics and tank design, e.g., kinetics that permit wider temperature swings are preferable. Achievable ACHX turndown also has implications for the MH selection. Multi-step MH reactions that occur over several enthalpies and temperatures are attractive because they reduce the average reaction enthalpy and increase the H2 capacity, but the varying enthalpy over the course of charging must be handled by the ACHX, which is designed to cool against ambient. The combination of turndown over the 3min refueling with turndown for daily and seasonal ambient temperature variation requires careful analysis. There is a risk of overcooling the coolant, which would then overcool the tank and slow down the hydriding reactions. It is difficult to make further progress without more data at scale for MH tanks, data that would allow an integrated simulation of the tank and refueling station that considers the recharging-chemistry energetics and kinetics, heat transfer for cooling, pressure drops incurred in the system (including refueling station), and coupling between heat transfer and kinetics; all are constrained by refueling-station considerations, especially achievable mass flow rates for the cooling fluid. A MH tank design that satisfies specific performance and cost targets must be developed to support further analysis. It is clear that there is a rich design space to consider for MH-based tanks for HFCEVs and it is also clear that the refueling-station design must be considered during the tank design because each has implications for and constrains the other. To our knowledge, MH tank design has not been constrained by refueling-station design in the published literature. At the very least, the refueling-station ACHX, the dispenser-vehicle interconnect, and the cooling fluid circulation power must be evaluated. It appears that these practical downstream considerations will affect the choice of the MH.

Acknowledgments This research was supported by the Fuel Cell Technologies Office of the U.S. Department of Energy’s Office of Energy Efficiency and Renewable Energy under Contract No. DE-AC0206CH11357. The authors thank Fred Joseck, Neha Rustagi, and Jesse Adam from the U.S. Department of Energy’s Fuel Cell

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Technologies Office for their guidance and support. The views and opinions of the authors expressed herein do not necessarily state or reflect those of the U.S. Government or any agency thereof. Neither the U.S. Government nor any agency thereof, nor any of their employees, makes any warranty, expressed or implied, or assumes any legal liability or responsibility for the accuracy, completeness, or usefulness of

any information, apparatus, product, or process disclosed, or represents that its use would not infringe privately owned rights.

Appendix. Station schematics

Fig. A.1 e Refueling station schematics. See also Ref. [7].

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