Experimental Testing of s-CO2 Regenerator for Use as a ...

EXPERIMENTAL TESTING OF S-CO2 REGENERATOR FOR USE AS A REPLACEMENT TO HIGH COST PRINTED CIRCUIT RECUPERATORS FOR USE IN S-CO2 RECOMPRESSION

BRAYTON CYCLE

Jacob F Hinze University of Wisconsin – Madison

Madison, WI, USA

Gregory F Nellis University of Wisconsin – Madison

Madison, WI, USA

Mark H Anderson University of Wisconsin – Madison

Madison, WI, USA

ABSTRACT Supercritical Carbon Dioxide (sCO2) power cycles have

the potential to deliver high efficiency at low cost. However, in

order for s-CO2 cycle to reach high efficiency, highly effective

recuperators are needed. These recuperative heat exchangers

must transfer heat at a rate that is substantially larger than the

heat transfer to the cycle itself and can therefore represent up to

24% of the total power block cost in a recompression Brayton

cycle [1]. Lower cost regenerators are proposed as a cost saving

alternative to high cost printed circuit recuperators. A

regenerator is a heat exchanger that alternately has hot and cold

fluid passing through it. During the first half of its cycle the hot

gas is passed over a storage media bed (stainless steel balls,

screens, or similar fill material) where thermal energy is stored.

During the next half of the cycle, cold fluid is passed through in

the opposite direction, extracting the thermal energy from the

bed. By operating a cycle with two (or more) regenerators,

where one is always in a hot to cold (HTC) blow and the other

in a cold to hot blow (CTH), a quasi-steady state can be

achieved in the cycle to allow continuous operation. A model of

the regenerator was created and used in place of a recuperator

in a model of a 10MW power plant. The thermal effectiveness

of the regenerator cycle was slightly lower than the recuperator

cycle, however the regenerator cycle had a saving of about 9.3

percent in the Levelized Cost of Energy (LCoE). A scale model

of the regenerator is under construction which will verify the

performance of the regenerator model. INTRODUCTION Supercritical carbon dioxide (sCO2) power cycles are very

simply a closed Brayton cycle operating with CO2 as the

working fluid. The main advantage to this cycle is the increase

in thermal efficiency over a typical Rankine cycle as shown in

Figure 1.

FIGURE 1 CYCLE EFFICIENCY VS TURBINE INLET

TEMPERATURE FOR VARIOUS WORKING FLUIDS FROM [1]

It is clear from Figure 1 that the other cycles do not achieve the

same efficiencies as the sCO2 cycle. Supercritical and

superheated steam operate at higher efficiencies than the sCO2

at low temperatures; however, once the turbine inlet

temperature is above 550°C a sCO2 cycle is more efficient. It is

desirable to operate at these high temperatures if the materials

will allow it since the cycle efficiency increases with

temperature. The helium Brayton cycle also allows the cycle to

operate at high temperature, however with a lower efficiency as

Proceedings of the ASME 2016 10th International Conference on Energy Sustainability ES2016

June 26-30, 2016, Charlotte, North Carolina

ES2016-59615

1 Copyright © 2016 by ASME

compared to sCO2. Additionally, helium is much more

expensive than CO2, meaning the seals in the system must be

very good to prevent any loss of helium. The main reason for

the high efficiency associated with the sCO2 cycle is the large

density change that occurs near the critical point. Pumping

power on a per mass basis is roughly equal to the difference in

pressure multiplied by the specific volume of the fluid. For a

Rankine cycle where liquid water is being pumped the energy

needed to pump the water back into the boiler is very small,

usually under 5 percent of the total output from the turbine.

This is in sharp contrast of an air Brayton cycle where the

compressor power can consume almost half of the work output

from the turbine. The high density of CO2 when it is

compressed means that the energy consumed by the compressor

is reduced to about 30 percent of the turbine power, increasing

the efficiency of the cycle. This along with the reduced

corrosion issues of CO2 as compared to steam suggest that

sCO2 cycles may be used for power generation in the future.

sCO2 power cycles have already found a market as a waste

heat recovery engine. In particular Echogen has developed an 8

MW waste heat recovery engine that is the size of

approximately 2 standard shipping containers. It operates with

waste heat at a temperature of 532°C and has a thermal

efficiency of 24 percent [2]. SCO2 offers several advantages

over other heat recovery engines including high efficiency, a

non-toxic/inexpensive working fluid, and compact size.

Alternative heat recovery engines such as organic Rankine

cycles use expensive working fluids and have lower

efficiencies.

As sCO2 technology becomes more mature there will be

many more opportunities for its use. The DOE’s goal for

concentrated solar power is to operate with a turbine inlet

temperature in excess of 700C [3]. Figure 1 shows that steam is

not a viable option at this temperature, and Helium has a

significantly lower efficiency. This leaves sCO2 as the obvious

choice. The size reduction of sCO2 turbo equipment is also

significant, a sCO2 turbine with an equal work output to a

steam turbine will be approximately 1/10th of the size [4],

resulting in lower capital costs, and thus a lower cost of

electricity.

sCO2 Cycle Configurations

The performance and cost of a sCO2 cycle can vary

drastically depending on the configuration of the cycle. A

simple Super Critical Brayton Cycle (SCBC) operates with

only a compressor, turbine, primary heat exchanger, precooling

heat exchanger, and recuperator as shown in Figure 2 (A). This

cycle requires a larger recuperator and the efficiency of the

cycle is limited because of the mismatch in capacitance rates of

the high and low temperature streams in the recuperator. When

CO2 is near the critical point, the specific heat capacity of the

fluid increases greatly which causes a mismatch in the

capacitance rate of the fluids. The Recompression Brayton

cycle (RCBC) seeks to balance the recuperators by adding a

second compressor and recuperator as shown in Figure 2 (B).

(A)

(B)

Figure 2 (A) SIMPLE CLOSED BRAYTON CYCLE (SCBC) AND (B) RECOMPRESSION CLOSED BRAYTON CYCLE

(RCBC) FROM [5]

The recompressor takes warm fluid leaving the low pressure

side of the low temperature recuperator and compresses it to

high pressure and returns it to the cycle after the low

temperature recuperator. Thus the mass flow rate of the low

pressure stream will be higher than the high pressure stream,

balancing the recuperator since the heat capacity of the low

temperature, high pressure CO2 is higher than the high

temperature, low pressure CO2 in the low temperature

recuperator. This allows the cycle to achieve a higher thermal

efficiency. Additionally there is thermodynamic advantage to

using a RCBC, since the fluid exiting the recompressor is at a

higher temperature meaning less heat is needed to reach the

turbine inlet temperature. The combination of balancing the

recuperators and adding the second compressor mean the

RCBC can operate at higher efficiencies than a SCBC for the

same sized recuperator [6].

One of the main barriers to widespread adoption of sCO2

power cycles is that it is untested at large scale and may have a

high cost. The power block is made up of the compressors,

turbine, primary heat exchanger, condenser, and recuperators.

There is substantial uncertainty in the cost of the recuperators

and it is estimated that this heat exchanger will account for

about 24 percent of the total power block costs [5]. In order for

sCO2 cycles to reach high efficiencies, they require highly

effective recuperators, usually greater than 90 percent.

Currently the state of the art recuperators are printed circuit

heat exchangers made by etching a pattern into many thin

layers of stainless steel and then diffusion welding them into a

block. These heat exchangers are very compact, offer high

effectiveness and low pressure drop, but they are very

expensive to build and are characterized by literally miles of


seals that all must be hermetic. The recuperators are the second

most expensive component behind only the primary heat

exchanger, as a result they are a good candidate for cost

reduction [5].

Regenerator Operation

Instead of a recuperator, a regenerator has been proposed

as a replacement device to accomplish the internal heat transfer

required by the cycle for high efficiency. A regenerator operates

by alternatively flowing a hot and cold fluid over an energy

storage bed. During the hot to cold (HTC) blow, hot fluid is

passed over the cold bed, and the thermal energy is stored in the

bed. Next during the cold to hot (CTH) blow, cold fluid is

passed over the bed and heat transferred to the fluid resulting in

it leaving at an elevated temperature. If this cycle is repeated

many times, the bed eventually reaches a quasi-steady state.

Since each regenerator canister can only have hot or cold fluid

passing through at any instance of time, at least two regenerator

canisters are needed to have the regenerator replace a

recuperator.

Since there is no need to have a hermetic seal between the

high pressure and low pressure streams, the regenerator

construction can be much simpler than for a recuperator. A

regenerator would consist of a tube packed with a filling

material (for example stainless steel balls) with headers welded

on and connected to a manifold of valves. Since there is no

complicated fabrication needed, it is assumed that the majority

of the costs for a regenerator will be the materials. Additionally,

since the sphere size in the bed can be made very small, the

surface area to volume ratio of a regenerator is much higher

than for a recuperator. More surface area means the regenerator

can have a higher effectiveness which results in a higher cycle

efficiency.

In order for regenerators to replace recuperators, they need

to provide an advantage in one or more categories of cost,

maintenance, reliability, or lifetime so that the replacement of

recuperators with regenerators creates a reduction in the LCoE.

The LCoE takes into account all of the capital costs associated

with the power block and spreads them out evenly to the

electricity produced by the cycle using the method outlined in

[7]. The LCoE calculations for this analysis were done for a

power plant with a 20MW thermal heat input and a turbine inlet

temperature of 700°C. The constant heat input takes into

account the difference in thermal efficiencies associated with

using regenerators compared to recuperators. A high efficiency

results in more power generation and thus more energy to

spread the capital costs over. The costs for the power block and

recuperators were obtained from [5] and the regenerator cost

was estimated using the mass of material needed and the cost of

this material inflated for manufacturing. A model was created

based on a RCBC operating with a 700°C turbine inlet

temperature and at pressures of 25 and 8 MPa. The recuperators

were specified to have a effectiveness of 90 percent which is

the maximum due to the manufacturing process where some of

the headers have to be in a cross flow configuration which

reduces the effectiveness [8]. The effectiveness of the

regenerators is therefore set to 95 percent since they can be

operated in a true counter flow configuration. To take into

account the losses in converting the shaft energy to electrical

energy it was assumed that the motors and generator had an

efficiency of 92.2 percent.. The thermal efficiency of the cycle,

the cost of the recuperator or regenerator, and the resulting

LCoE are summarized in Table 1.

TABLE 1 EFFICIENCY, RECUPERATOR COSTS, AND LCoE FOR A RECUPERATOR AND REGENERATOR OPERATING

AT IDENTICAL CONDITIONS Thermal

Efficiency (%)

LCOE

($/kWh)

Recuperator/

regenerator

cost (k$)

Recuperators 40.1 0.02087 2250

Regenerators 38.1 0.01893 971

The cycle operating with a regenerator has a similar thermal

efficiency (slightly smaller due to the carryover effect) and a

9.3 percent reduction in LCoE due to the much lower cost of

the regenerator. This simple calculation of costs savings shows

that regenerators have the potential to reduce the costs of the

sCO2 power cycle.

REGENERATOR DESIGN While regenerators do not see widespread use in the power

industry currently, regenerators have been used successfully for

many years for HVAC and cryogenic applications. The design

of a regenerator for use in power cycle will modify models

already used in these industries. In [9] an effectiveness-NTU-

Cm method is defined for regenerators. This method is very

similar to the effectiveness-NTU method for heat exchangers,

but this model adds an additional parameter, the matrix capacity

ratio defined by equation (1).

0 min

b bm

m cC

P C (1)

where mb is the mass of the packed bed, cb is the specific heat

capacity of the bed, P0 is the switching period, and minC is the

minimum capacitance rate of the fluids passing through the

bed. The capacitance rate of each stream is defined by equation

(2).

,p avgC c m (2)

where cp,avg is the average specific heat capacity through the

regenerator of the stream and m is the mass flow of the fluid. It

is important to understand the effect that changing the matrix

capacity ratio has on the effectiveness of the regenerator. A high

matrix capacity ratio means that the bed can hold much more

energy than the fluid can impart, and temperature of the packed

bed will change very little during the cycle. In the extreme case

where the matric capacity ratio approaches infinite, the

regenerator will behave in exactly the same way as a

recuperator because, as in a recuperator, the bed temperature

will not change with time. As the matrix heat capacity is

reduced, the effectiveness of the regenerator also decreases.

This is because the temperature of the bed changes more


dramatically during each cycle as it absorbs and releases

energy. After the HTC blow, the temperature at the hot end of

the regenerator will be approximately equal to the temperature

of the hot fluid coming in. During the CTH blow, cold fluid

will start flowing through the bed in the opposite direction and

at the beginning the exit temperature will be very close to the

temperature of the hot fluid, however as the cycle continues the

temperature will decrease since the temperature of the bed on

the hot side is being reduced by the cold fluid passing over it.

The temperature of the cold fluid leaving the regenerator during

the HTC blow is shown in Figure 3.

FIGURE 3 TEMPERATURE OF FLUID LEACING HOT SIDE OF REGENERAOTR DURING THE HOT TO COLD BLOW

The temperature that really matters is the average fluid exit

temperature, and as the matrix capacity ratio gets lower, the exit

temperature will fall further. When averaged over the mass flow

out of the regenerator the average exit temperature will be

lower and as a result the effectiveness of the regenerator will be

lower.

The number of transfer units, NTU, also factors

heavily into the effectiveness of the regenerator. Much like a

recuperator, a higher NTU will result in higher effectiveness.

The NTU is function of the minimum capacitance rate and the

conductance defined by equation (3).

1

1 1

H s C s

UA

hc A hc A

(3)

where hcH is the average heat transfer coefficient for the HTC

blow, hcC is for the CTH blow, and As is the surface area that is

exposed to the fluid. NTU is defined according to equation (4).

min

UANTU

C (4)

It is much easier to have a large NTU in a regenerator as

compared to a recuperator because the specific surface area of a

bed of spheres is much higher than a printed circuit heat

exchanger and the passages are quite small leading to high heat

transfer coefficients. A correction needs to be added to the

matrix capacity ratio and NTU since the capacitance rate on

each side of the regenerator are not the same. The matrix

capacity ratio is adjusted according to equation (5).

,

2

1

R mm e

R

C CC

C

(5)

The NTU is corrected according to equation (6).

2

1

Re

R

C NTUNTU

C

(6)

In both equations CR is the capacitance ratio defined by

equation (7).

min

max

R

CC

C (7)

The model takes the corrected NTU and matrix capacity ratio

and determines the effectiveness [9].

The carryover mass is an unavoidable loss associated with

the operation of a regenerator that is not present with a

recuperator. The carryover mass is the high pressure fluid that

is stuck in the regenerator when it is switched from the CTH

blow to the HTC blow. Since the pressure difference is large

between the CTH and HTC blow, the specific volume of the

fluid differs greatly. This means the amount of mass that enters

the regenerator at high pressure and then returns to the

compressor without going through the turbine during the

switching process can be significant, up 25 percent of the total

mass that passes through the regenerator during the flow

processes depending on the conditions. This carryover mass is a

direct reduction to the cycle efficiency since the energy used to

compress it is not recovered in the turbine. Having a large

carryover mass will result in a lower thermal efficiency which

in turn reduces the LCoE; therefore, it is important to minimize

the carryover as much as possible. The carryover mass is

calculated by equation (8).

2

0(1 )( ( , ) ( , ))

4

L

c v H L

πm D e rho T P rho T P dL

(8)

where D is the inner diameter of the regenerator, ev is the void

volume fraction of the packed bed (0.37 for a bed of spheres),

and rho(T,PH) and rho(T,PL) is the density of the fluid at

location L in the canister at the point that the flow process is

complete for the high pressure and low pressure cases. As a

rough estimation it assumed that the temperature of the fluid in

the canister varies linearly from the hot inlet temperature to the

cold inlet temperature. Since there are two regenerator beds,

this mass of fluid is returned to the compressor twice every

period. This carryover mass is quantified as a mass flow rate

according to equation (9) and treated as a leakage in the system.


0

2 cc

mm

P (9)

The model also calculates a pressure drop based on the size

of the packed bed of spheres, and the average properties of the

fluid using a correlation for flow in a packed bed [10]. The

effectiveness-NTU-Cm model expects an input for the size of

the regenerator, however it is desirable to specify the

effectiveness of the regenerator and have the model determine

the size. To accomplish this the iterative equation solving in

EES is used [10]. The user specifies the effectiveness, sphere

size, switching period, and pressure drop, and the model

iteratively solves for the diameter, length, and carryover mass

for a regenerator meeting the desired operating parameters.

Design Goals

The goal for designing the regenerator is to have the largest

possible reduction in LCoE due to the use of regenerators rather

than recuperators. The main reduction in LCoE comes from the

reduced capital costs of the regenerators as compared to the

recuperators. To meet this goal the regenerator should be kept

as small as possible. There are two ways of having a high

effectiveness regenerator, the first is having a large matrix

capacity ratio, and the second is having a high NTU. Equation

(1) shows that to keep the same Cm for a given minimum

capacitance rate the mass of the bed must be increased if the

switching time is increased. Therefore, the switching time

needs to be kept as short as possible. For a set switching time

the only way to increase matrix capacity ratio is to increase the

mass of bed which will increase the cost. So instead of having a

large matrix capacity ratio the more economical method is to

have a large NTU. The NTU is based on the UA of the cycle

which can be increased by simply making the size of the

spheres smaller. This increases the UA two ways, first by

increasing the surface area for heat transfer, and second by

increasing the heat transfer coefficients by decreasing the

hydraulic diameter. The disadvantage of this is that the pressure

drop also increases with a smaller hydraulic diameter.

Limitations for Design Parameters

Several limits need to be placed on what is possible for

these parameters. The first is to constrain the pressure drop, since

increasing the pressure drop directly decreases the cycle

efficiency. Therefore, the maximum pressure drop was set to 1

percent of the absolute pressure of the cycle or 250 kPa. The

pressure drop is controlled by changing the aspect ratio

(length/diameter) of the regenerator, a small aspect ratio will

result in a smaller pressure drop. The switching time cannot be

too small since the valves need time to respond. The valve

response time was estimated to be 3 seconds to fully open and

close and therefore so the smallest allowable flow period is set

to 30 seconds. The spheres in the packed bed can only be made

so small, it was assumed that the smallest possible size sphere is

1mm. Stress analysis was conducted on standard pipe sizes to

determine what size pipe can be used to construct the

regenerator. The largest pipe size that would be able to tolerate

the temperature and pressure in the regenerator is a 24 inch

schedule 160 pipe. This is smaller than the diameter needed for

the regenerator so it is assumed that many tubes will be plumbed

in parallel to obtain the proper mass and pressure drop. A larger

diameter pipe allows for fewer pipes to be used, lowering the

total manufacturing costs of the regenerator. The material for the

outer tube is 316 and the packed bed is 304. For the optimized

designed conditions the dimensions of the high and low

temperature regenerators are listed in Table 2.

TABLE 2 REGENERATOR DIMENSIONS FOR 10MW POWER PLANT

Low Temp Regenerator

High Temp Regenerator

Length (m) 1.34 (52.8 in) 0.86 (33.9 in)

Reqd. Total Frontal Area (m2)

1.27 (13.7 ft2) 1.37 (14.7 ft2)

Regenerator Volume (m3)

1.70 (60.0 ft3) 1.17 (41.3 ft3)

Number of 24 inch Pipes Needed

14 16

Numerical Validation

In order to verify that the relatively simple effectiveness-

NTU-Cm model with empirical correction factors was providing

correct results, a MATLAB [11] numerical model was created

to check the solution based on the numerical one-dimensional

regenerator model outlined in [12]. This model is provided with

the dimensions of the regenerator, the temperature and mass

flow rate of the fluid entering the regenerator at each side, and

average heat transfer coefficient for each blow of the

regenerator. The model uses matrix decomposition to calculate

the temperature of the fluid throughout the cycle using the

average capacitance rate of the fluid. Then using successive

substitution it calculates the heat transfer rate based on the heat

transfer rate based on the actual properties at each node of the

regenerator. The numerical model predicts a heat transfer rate

that is approximately 5 percent larger for the low temperature

regenerator, and 2.5 percent larger for the high temperature

regenerator. The numerical model shows good agreement with

the effectiveness-NTU-Cm model and therefore provides

confidence in the results it predicts.

REGENERATOR IMPROVEMENTS The main area for improvement of the regenerator system

is decreasing the carryover mass. One of the simplest ways of

doing this is decreasing the void volume fraction of the

regenerators. The void volume fraction is the fraction of

volume in the regenerator that is occupied by the working fluid.

For a packed bed of spheres the normal void volume fraction is

0.37, if this reduced to 0.26 the theoretical limit for packed

spheres [13], the thermal efficiency is increased to 39.6 percent.

This could be accomplished by using a mix of large and small

diameter balls, where the small balls nest inside of the voids

between the large diameter balls.

Another way to decrease the carryover mass is by using the

high pressure gas in one regenerator to partially fill the low

pressure regenerator rather than allowing this mass to drain

directly back to the compressor. If a valve is introduced that


allowed the two regenerators to exchange mass with each other

before opening the valves to the system lines, the high pressure

fluid that would have returned directly to the compressor

instead comes to equilibrium with the low pressure fluid in the

other regenerator. Then, once the pressure has equalized, the

valves are opened, and normal operation is resumed. Operating

the valves in this fashion would reduce the carryover mass by a

factor of 2 and increase the thermal efficiency from 38.1

percent to 40.0 percent. This valve operation is illustrated in

Figure 4.

Regenerator

Regenerator

High Pressure

Inlet

Low Pressure

Exit

High Pressure

Exit

Low Pressure

Inlet

Bypass Valve

Regenerator

Regenerator

High Pressure

Inlet

Low Pressure

Exit

High Pressure

Exit

Low Pressure

Inlet

Bypass Valve

Regenerator

Regenerator

High Pressure

Inlet

Low Pressure

Exit

High Pressure

Exit

Low Pressure

Inlet

Bypass Valve

HP

HP

LP

LP

EP

EP

FIGURE 4 BYPASS VALVE OPERATION, HP = HIGH

PRESSURE, LP = LOW PRESSURE, EP = EQUILIBRIUM PRESSURE

As Figure 4 shows the bypass valves requires that all

valves to each regenerator be closed except for the bypass

valve. This means that there will be nowhere for the high

pressure fluid exiting the compressor to go. To eliminate this

problem, the regenerator is broken into multiple beds so that

continuous flow can be maintained. This means that while one

pair of beds are pressurizing/depressurizing the other beds are

still operating normally. This also serves a secondary purpose

of being able to operate more effectively at part load

conditions. If all of the beds were used under part load

conditions the regenerators would be much larger than needed

to obtain high effectiveness. That extra mass would result in

larger carryover relative to the mass flow through the system

reducing the thermal efficiency. Instead unneeded regenerators

can be removed and the cycle can always be operating with

essentially an optimally sized regenerator.

VALVE CHOICE AND TESTING

Almost as important as the regenerators themselves are the

valves the will direct the flow of the high temperature and

pressure CO2 into the proper regenerator. Not only is the

temperature in the valves very high (560ºC for a 700ºC

turbine), it is also not constant during each cycle. Figure 3

shows the temperature variation that the valve will experience

every 15 seconds as the cycle is running. Rapid temperature

changes cause thermal stresses in the valve body which can

result in leakage or premature breakage of the valves. Ideally a

valve would last for the lifecycle of the power plant (20 years)

but that may not be feasible in such a hostile environment. The

very shortest limit for the valve is one year which is how often

most plants shutdown for maintenance. The valves also need to

be able to open and close quickly so that regenerators can

operate correctly. The valve opening/closing time as well as

lifespan while determine what switching period can be used in

the cycle.

There are several types of valves that can withstand the

pressure and temperature experienced in the cycle. The first and

simplest option would be to use an open/closed globe valve.

Each end of the regenerator would have two valves, one that

goes to the high pressure side of the loop, and the other to the

low pressure side, as shown in Figure 4. On the CTH blow, the

valves on each end of the regenerator connected to the high

pressure side of the system would open up, and the high

pressure fluid exiting from the compressor would remove heat

from the bed, in the next half of the cycle the regenerator would

be connected to the low pressure side of the cycle and heat

would instead be stored in the bed. The main disadvantage of

this option is the number of valves needed, even in the simplest

configuration with two bundles of regenerators, 8 valves are

needed, and if the regenerator is broken into smaller bundles

then 4 valves are needed for each bundle.

The next simplest valve would be to use a 3-way valve.

Instead of needing two valves at each end of the regenerator

only one valve is needed since the same valve can connect to

both the high pressure and low pressure connections. The valve

is switched from one connection to the other by rotating the

valve body. This valve would operate in the same way to the

globe valve and would cut the number of valves and actuators

needed in half.

The most complicated valve would be a rotary ball valve.

This valve would have a central mandrel which rotates and

directs flow into ports cut into the body of the valve. Each

regenerator bundle would have its own port and each rotation

would alternatively connect high the and low pressures flows to

the regenerator bundle. Depending on the size of the valve,

many different regenerator bundles could be connected to the

same valve. This means that for a 10MW plant only four valves


are needed. Of course the issues with sealing, wear, and

maintenance are much greater for rotary valves than for the

other two types because of its complexity.

Choosing the right valve depends on the lifetime costs of

the valve systems. Information is being gathered in

collaboration with a major valve manufacturer in order to

compare the capital and maintenance costs for each type of

valve. The valve system with the lowest overall costs while

maintaining an acceptable level of reliability will be chosen for

use.

SCALE EXPERIMENTAL TESTING To validate the effectiveness-NTU-Cm model and

MATLAB numerical model, a scaled regenerator will be tested

at the same temperature expected in the full scale plant but with

a heat transfer rate of 10kW. However, for the initial test the

pressure will be reduced to 16 MPa due to the limits of the

available compressor. In addition to verifying the model it will

also provide an opportunity to test more practical

considerations of a regenerator system such as the valves,

control scheme, construction, and reliability. The performance

of the small scale system will determine the validity of the

model, the feasibility of using regenerators for power

production, and determine where to focus further research

efforts. The test regenerator will be constructed so that it

operates at the same temperature, pressure, and effectiveness as

the high temperature regenerator in the 10MW plant since this

is the more extreme environment.

To ensure that the results from the 10kW test can be scaled

up to the regenerators for the 10MW power plant, there are two

loss mechanisms that need to be matched from the large

regenerator to the test regenerator. The first is energy storage in

the wall of the regenerator. A simple one dimensional transient

model of the wall of the regenerator was constructed to

determine the amount of energy that would be stored in the wall

every cycle. For the 10MW high temperature regenerator the

amount of energy stored in the wall compared to the energy

stored in the bed was small, under one percent. The energy

stored in the wall doesn’t necessarily hurt the performance of

the regenerator since it just act as an extension of the packed

bed. The second more important inefficiency is axial

conduction along the wall of the regenerator. Axial conduction

takes high temperature heat and transfers it to the low

temperature fluid, this means more energy is leaving the

regenerator on the HTC blow reducing the effectiveness of the

regenerator. As a result it is much more important to match the

axial conduction from the 10MW regenerator to the 10kW

regenerator. The temperature difference between the hot and

cold ends is the same for both the full and subscale cases;

therefore, the amount of conduction will depend on the

resistance to conduction through the wall of the regenerator. As

schedule 160 pipe is scaled down to smaller diameter, the ratio

of internal area to wall area gets smaller. So, per unit length

there is less area for the bed and more area for conduction. The

aspect ratio of the 10MW high temperature regenerator is very

small, if this is scaled down to the 10kW test regenerator the

axial conduction is quite high. Therefore, the aspect ratio itself

is made much higher in the 10kW device. In order to have a fair

comparison of the axial conduction, the percentage of heat

conducted axially is compared to the total ineffectiveness of the

regenerator, as shown in equation (10).

max

100%axialaxial

act

QPer

Q Q

(10)

where Qmax is the maximum amount of heat transfer the

regenerator could have, Qact is the actual heat transfer of the

regenerator, and Qaxial is the heat conducted axially down the

wall of the pipe. The axial conduction was calculated using the

resistance of the pipe wall as shown in equation (11).

,( )

c pipe

axial hot cold

pipe

k AQ T T

L (11)

For the high temperature regenerator of the 10MW power

plant this was calculated to be only one percent of the total

ineffectiveness. However if the same aspect ratio of about 1

was used for the 10kW regenerator, the axial conduction would

be approximately 15 to 20 percent of the ineffectiveness. In

order to minimize the conduction the design for the regenerator

was stretched to have an aspect ratio of about 9. The axial

conduction percentage is about 2 percent of the ineffectiveness

in the regenerator which is very small just like its 10MW

counterpart.

Currently the compressor in the experimental test facility is

only rated to 16 MPa, so instead of schedule 160 pipe, schedule

80 pipe can be used which also helps reduce the energy stored

in the wall and axial conduction. The specifications for the test

regenerator are listed in Table 3.

TABLE 3 DIMESIONS OF TEST REGENERATOR

Effectiveness 0.95

Switching Time (s) 60

Length (cm) 45.1 (17.7 in)

Diameter (cm) 4.93 (1.94 in)

Wall thickness (mm) 5.5 (0.216 in)

Sphere size (mm) 3 (0.125 in)

Wall Storage Percent 18

Axial Conduction Percent

2.1

The switching time has also been increased to 60

seconds out of concern for the large temperature change that the

fittings have to endure. By increasing the cycle time the

temperature transients are reduced which should increase the

lifetime of the system components.

Using the design in Table 3 a set of regenerators has

been constructed for testing. One of the most important factors

determining how well the regenerator will operate is the mass

of the packed bed material. It is important therefore to have the

tightest packing possible to closely march the model. Packing

was accomplished by incrementally adding the stainless steel

balls to the tube and vibrating them into place. At three points

in each regenerator a thermocouple was inserted into the


packed bed through a wall of the pipe. This allows for in-situ

measurement of the packed bed temperature which will help

determine the accuracy of the model. After the regenerator was

filled, the actual mass used to fill was compared to the

theoretical mass calculated from the model, that packing

percentage with very good, about 98 percent for both

regenerators.

The spheres are held in place by stainless steel screens

welded onto the end of the pipe. This was accomplished by

clamping the screen in between the pipe and the end cap, and

welding the endcap into place. When the regenerator was filled

to the top with the balls, the other screen was placed on top of

the pipe and covered with the end cap. The end cap was then

welded into place taking care to not melt the screen. On each

end of the regenerator there will be two thermocouples

measuring the temperature, and one pressure tap, which will

allow measurement of both the absolute pressure at the inlet of

the regenerator and the differential pressure across the

regenerator. The constructed regenerator is shown in Figure 5.

FIGURE 5 CONSTRUCTED REGENERATORS FOR 10kW

TEST FACILITY

EXPERIMENTAL TEST FACILITY In addition to pressure and temperature measurements in

the regenerator, the mass flow rate at the cold end of each

regenerator will be measured using an orifice flow meter. The

pressure drop over the orifice will be calibrated throughout the

expected range of operating temperature and pressures using a

Coriolis flow meter. It is necessary to measure the mass flow

near the regenerator to get an accurate measurement of the

mass flow rate of the fluid entering the regenerator. If mass

flow meter is located far away from the regenerator the peaks

and troughs of flow will be eliminated by the buffer volumes

and it will be difficult to determine the actual mass flow rate

experienced by the regenerator as a function of time. The

Coriolis flow meters are not able to withstand the high pressure

and temperatures that the regenerator will experience, so they

must be located before and after the compressor to measure the

mass flow rate through the compressor. The cycle diagram is

shown in Figure 6.

The compressor available is a hydraulically driven CO2

compressor capable of providing CO2 at 16 MPa. In order to

operate at 25 MPa some modifications will eventually be

needed. The compressor will need to be changed to have a

smaller bore so it can reach a high pressure, or an additional

pump will be needed to boast the pressure. Initial testing will

occur at 16 MPa since the properties of sCO2 do not change

much between 16 MPa and 25 MPa so if the model matches the

experimental results at 16 MPa, it should also work at 25 MPa.

The temperatures at states 2 and 4 are controlled using the

preheater and main heater respectively. The temperature at the

inlet to the compressor is controlled by the water flow rate used

in the compressor. The pressure differential and mass flow rate

through the system will be controlled using the bypass valve

and expansion valves. The system will measure the temperature

at states 3 and 5. Along with flow rate measurements these

temperatures can be used to calculate the effectiveness of the

regenerator. Not shown on the diagram are the pressure

differential across the regenerator which will be used to verify

the analytical model.

CONCLUSIONS The simple effectiveness-NTU-Cm model indicated

that a regenerator optimized to reduce LCoE would

have a Cm near one, a high NTU, and a short switching

time

More sophisticated manufacturing and operating

techinques can increase the effectiveness of

regenerators and provide an even larger savings to

LCoE

FIGURE 6 PROCESS DIAGRAM OF EXPERIMENTAL TEST FACILITY


The valves needed to control this system pose a

significant challenge, however valves are available

that offer long life at high temperature which will need

to be verified through experimental testing

The 10kW test facility will allow verification of the

regenerator model and provide information about the

practical aspects of operating a high temperature

regenerator

sCO2 power cycles has the potential to provide

electricity at much greater efficiency than current Rankine

cycle technologies due its ability to operate at much high

temperatures. However the high degree of recuperation

necessary means that the printed circuit heat exchangers

necessary are very costly [5]. An initial analysis conducted

compared the costs of a regenerator to a printed circuit heat

exchanger and found that the regenerator would likely cost

half as much as a recuperator. The cost reduction resulted

in a 9.3 percent reduction in LCoE. Regenerators offer the

opportunity to build and operate high temperature sCO2

cycles at a significantly reduced cost compared to a

recuperated RCBC.

NOMENCLATURE Ac,pipe = Cross sectional area for conduction in pipe

As = Surface area of packed bed

C = Capacity rate

cb = Specific heat capacity of packed bed

Cm = Matrix capacity ratio

cm,e = Equivalent matrix capacity ratio

maxC = Maximum capacity rate

minC = Minimum capacity rate

cp,avg = Averages specific heat capacity

CR = Capacitance ratio

CTH = Cold to Hot

D = Diameter of regenerator

ev = Void volume fraction

hcC = Heat transfer coefficient on cold side of regenerator

hcH = Heat transfer coefficient on hot side of regenerator

HTC = Hot to Cold

LCoE = Levelized Cost of Electricity

Lpipe = Length of regenerator pipe

k = Conductivity of shell material

NTU = Number of Transfer Units

NTUe = Equivalent Number of Transfer Units

m = Mass flow rate

mb = Mass of packed bed

mc = Carryover mass

mc = Carryover mass flow rate

P0 = Switching period

Peraxial = Percentage of axial conduction

PH = High pressure in regenerator

PL = Low pressure in regenerator

RCBC = Recompression Brayton cycle

SCBC = Supercritical carbon dioxide Brayton cycle

Qact = Actual heat transfer in regenerator

Qaxial = Axial conduction in regenerator shell

Qmax = Maximum possible heat transfer in regenerator

Tcold = Low temperature of regenerator

Thot = High temperature of regenerator

UA = Conductance

ACKNOWLEDGMENTS This material is based upon work supported by the U.S.

Department of Energy under Award Number DE-EE0007120.

This report was prepared as an account of work sponsored

by an agency of the United States Government. Neither the

United States Government nor any agency thereof, nor any of

their employees, makes any warranty, express 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. Reference herein to any

specific commercial product, process, or service by trade name,

trademark, manufacturer, or otherwise does not necessarily

constitute or imply its endorsement, recommendation, or

favoring by the United States Government or any agency

thereof. The views and opinions of authors expressed herein do

not necessarily state or reflect those of the United States

Government or any agency thereof

REFERENCES [1] M. J. Driscoll, “Supercritical CO 2 Plant Cost

Assessment,” 2004.

[2] Echogen, “EPS100 Heat Recovery Solution,” 2015.

[Online]. Available: http://www.echogen.com/our-

solution/product-series/eps100/.

[3] DOE and EERE, “Concentrating Solar Power:

Advanced Projects Offering Low LCOE Opportunities,”

2014.

[4] M. Persichilli, A. Kacludis, E. Zdankiewicz, and T. Held,

“Supercritical CO 2 Power Cycle Developments and

Commercialization: Why sCO 2 can Displace Steam Ste

am,” Power-Gen India Cent. Asia, pp. 1–15, 2012.

[5] C. K. Ho, M. Carlson, P. Garg, and P. Kumar, “Cost and

Performance Tradeoffs of Alternative Solar-Driven S-

CO2 Brayton Cycle Configurations,” 2015.

[6] J. J. Dyreby, “Modeling the Supercritical Carbon

Dioxide Brayton Cycle with Recompression,” 2014.

[7] W. Short, D. J. Packey, and T. Holt, “A Manual for the

Economic Evaluation of Energy Efficiency and

Renewable Energy Technologies,” 1995.

[8] T. Held, “Performance & cost targets for sCO 2 heat

exchangers,” in NETL-EPRI Workshop on Heat

Exchangers for Supercritical CO2 Power Cycles, 2015.

[9] R. Barron and G. Nellis, Cryogenic Heat Transfer, 2nd

ed. Taylor and Francis, 2016.

[10] F-Chart, “EES Engineering Equation Solver.” 2015.

[11] MathWorks, “Matlab2015a.” 2015.

[12] G. Nellis and S. A. Klein, Heat Transfer. New York:

Cambridge University Press, 2009.

[13] W. Zhang, K. E. Thompson, A. H. Reed, and L. Beenken,

“Relationship between packing structure and porosity in

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