C. Gould, Fluor, Aliso Viejo, California; and R. JOYNER and A. DESHMUKH, Fluor, London, UK
As the world continues along its energy transition trajectory, demand for blue hydrogen (H2) and reduced carbon emissions will make high-purity carbon dioxide (CO2) streams available in many process plants. This article investigates how these streams could be utilized to add value to projects.
Part 1 of this article explores the potential use of supercritical CO2 as a refrigerant, replacing conventional refrigerants using hydrocarbons or hydrofluorocarbons (HFCs) to provide chilling duty on large energy and chemicals projects. The global warming potential (GWP) of CO2 is significantly lower than commonly used HFC refrigerants; therefore, the use of CO2 as a replacement for many HFC refrigerants would present a reduction in greenhouse gas (GHG) emissions when measured in equivalent mass of CO2.
This article considers how the operating conditions of CO2 refrigeration systems can be optimized to minimize energy consumption and improve efficiency. It also discusses the use of process simulation tools to compare a CO2 refrigerant system with a conventional propane refrigeration system regarding capital and operating costs (CAPEX/OPEX), operability, safety, and the potential for integration with carbon capture and sequestration (CCS) technology.
The history of CO2 as a refrigerant. The use of CO2 as a refrigerant dates back to the 19th century, with the first patent for a refrigerator using R-744 (CO2) being granted in 1850.1 The popularity of CO2 as a refrigerant increased in the early 20th century due to its lower flammability and toxicity vs. the two most common alternative refrigerants at the time: ammonia (NH3) and sulfur dioxide (SO2). By the 1930s, 80% of marine vessels used CO2 for refrigeration and comfort air cooling,1 and it was used universally by British ships into the 1940s.2 However, following the introduction of refrigerants using chlorofluorocarbon (CFC) in the early 1930s, CO2 began to be phased out and replaced because of the high versatility and reduced system complexity offered by these new synthetic refrigerants. By the 1950s, synthetic refrigerants had come to dominate industrial refrigeration applications.3
However, interest in the applications of CO2 as a refrigerant has persisted, particularly following the phase-out of ozone-depleting CFC refrigerants in the 1980s and 1990s. CO2 has been effectively applied in automotive air conditioning and commercial refrigeration systems,3 and its use in the low-temperature stage of industrial cascade refrigeration systems was common by the 2000s.4
As further restrictions on refrigerants with a high GWP came into effect (e.g., the 2014 EU F-Gas Regulation5), the need for alternative refrigerants intensified. In 2008, there were fewer than 150 transcritical CO2 refrigeration systems across Europe. By 2020, there were more than 29,000, predominantly in commercial refrigeration applications.6 However, CO2 refrigerant has yet to take hold in large energy and chemicals plants, where the use of synthetic refrigerants or hydrocarbons, such as propane and butane, is still most common.
GWP. CO2 is forecast to contribute approximately 78% of the equivalent GHG emissions between 2013 and 2050.7 However, the high impact of CO2 is predominantly due to its abundance, rather than its potency, as a GHG. CO2’s GWP is significantly lower than commonly used HFC refrigerants, as shown in TABLE 1. Overall, HFC emissions are increasing at a rate of 8%/yr, and emissions are projected to rise to 7%/yr–19%/yr of global equivalent CO2 emissions by 2050.8 In the Kigali Amendment to the Montreal Protocol, 197 countries committed to cut the production and consumption of HFCs by more than 80% over the next 30 yr.9 The use of CO2 as a replacement for any of the HFC refrigerants listed in TABLE 1 would therefore present a significant reduction in GHG emissions when measured in terms of GWP, even if the CO2 used was eventually released into the atmosphere.
The GWP of alternative hydrocarbon refrigerants (e.g., propane and butane) is lower than CO2; therefore, to realize the benefits in terms of reducing GHG emissions, the CO2 used as a refrigerant would have to be eventually sequestered or otherwise utilized in a manner that did not result in it being released into the atmosphere. As this article will detail, the potential exists to integrate a CO2 refrigeration system with CCS plants.
CO2 refrigerant system. The basic refrigeration system considered in this article is a simple closed-loop refrigeration cycle (FIG. 1). This configuration is fundamentally the same, regardless of which refrigerant is used. The principal difference is that, with a conventional refrigerant such as propane, the compressor aftercooler is a total condenser; whereas in the CO2 configuration, it is a cooler with no phase change. This will be discussed further in the following section.
The refrigerant is a saturated vapor at the suction of the compressor. The compressor increases the pressure of the refrigerant, which is then condensed or cooled and routed to the accumulator. The accumulator acts as a buffer vessel before the refrigerant moves to the chiller. The cool refrigerant from the accumulator is flashed through the Joule-Thomson (J-T) valve to a low pressure, resulting in a cold two-phase stream, which is routed to the chiller. The liquid portion of the refrigerant from the J-T valve is fully vaporized in the chiller, and the resulting vapor is routed to the suction of the compressor via a knockout drum, completing the cycle. The chilling duty is provided by the latent heat of the liquid refrigerant as it vaporizes in the chiller. The gas portion has minimal impact on the heat duty.
The impact of operating pressure. In a conventional, sub-critical refrigeration loop, the compressor’s discharge pressure is determined by the vapor pressure of the refrigerant at the condensing temperature, typically based on the minimum approach to the cooling medium. The refrigerant would therefore be a saturated liquid when it is supplied to the J-T valve. The refrigerant may be a subcooled liquid at the J-T valve inlet should the system include an economizer.
However, the critical temperature of CO2 is 31.1°C. Consequently, liquid CO2 cannot be formed above this temperature regardless of the pressure. Many plants utilize atmospheric air as a cooling medium. Given that an approach in the region of 10°C to the maximum design ambient air temperature is typically required, this means that condensing CO2 at 31°C is infeasible by air cooling in all but the coldest of climates. As the formation of a liquid CO2 refrigerant stream is infeasible, the alternative is to produce a supercritical CO2 stream at conditions such that a low-temperature, two-phase stream is produced when it is flashed through the J-T valve.
Given that the CO2 stream will be supercritical at the compressor discharge conditions, it will not condense through the compressor aftercooler, as there is no observable phase change between liquid and supercritical or gas and supercritical phases. Without the limitation of operating at the vapor pressure of the refrigerant, there is a greater degree of freedom in the selection of the compressor discharge pressure. This prompts the following question: Is there an optimum pressure at which to operate the discharge of the compressor?
FIG. 2 shows a pressure-enthalpy diagram of CO2. As an example, consider a plant that requires a process stream to be chilled to 5°C—and thus a CO2 refrigerant stream at 0°C—to allow a reasonable temperature approach in the chiller. Air cooling is available with a design ambient air temperature of 35°C. A refrigerant-side outlet temperature of 45°C from the air cooler is therefore considered. The CO2 refrigerant will leave the chiller shell as a saturated vapor. The boiling refrigerant temperature will be maintained by controlling the compressor’s suction pressure. Conditions in the shell are plotted in FIG. 2 at the intersection of the dewpoint curve with the 0°C isotherm, which occurs at 33.5 barg.
The compressor discharge conditions can be found through process simulation. The line representing the path through the compressor in FIG. 2 corresponds to a compressor adiabatic efficiency of 75%. The gradient of the line is a function of the compressor’s efficiency. The steeper the line, the higher the efficiency, thus, the lower the discharge temperature. Considering constant efficiency, any compressor discharge pressure will fall on this line. Four points have been shown in FIG. 2, corresponding to compressor discharge pressures of 100 barg, 120 barg, 140 barg and 160 barg.
A horizontal line can be drawn from the compressor discharge until it intersects with the 45°C isotherm. This represents the compressor aftercooler, with the end of the line representing the inlet to the J-T valve. Given that the J-T valve is adiabatic, the path through the valve can be drawn vertically down from the point representing the outlet of the aftercooler until it intersects the 0°C isotherm. This point is in the two-phase region for all four of the cycles drawn in FIG. 2.
The chilling duty is provided by the latent heat of the CO2 as it vaporizes; therefore, for all four cases, the liquid flowrate will be the same. However, the vapor fraction at the outlet of the J-T valve will be different. As the vapor fraction increases, the vapor flow associated with the liquid also increases. The consequence of this is an increased circulation rate; hence, increased compressor size and power consumption. Alternatively, consider that for all four cycles, the J-T valve outlet is at the same temperature and pressure but with different molar enthalpies. The lower the compressor discharge pressure, the higher the molar enthalpy at the outlet of the J-T valve. The chiller duty is equal to the difference in molar enthalpy between the inlet and outlet of the exchanger multiplied by the molar flowrate of the refrigerant. Therefore, a higher inlet molar enthalpy results in a lower molar enthalpy change across the exchanger; thus, a higher flowrate is required to achieve the same duty.
Given that increasing the compressor discharge pressure will reduce the circulation rate of the refrigerant required to meet the same chilling duty, this should reduce the power demand for the compressor. However, increasing the compression ratio of the compressor should conversely result in increased power demand. Another notable aspect of the cycles plotted in FIG. 2 is that there are diminishing returns in reducing the enthalpy at the inlet to the chiller as the compressor discharge pressure is increased—i.e., the gap between the J-T valve outlet points for the 100 barg and 120 barg cycles is larger than from 120 barg to 140 barg and so on. This is because the 45°C isotherm becomes increasingly steep as the pressure increases beyond ~90 barg. These factors imply that there may be an optimum compressor discharge pressure that minimizes the required power of the compressor.
This relationship is revealed by plotting compressor power vs. compressor discharge pressure (FIG. 3) in the range of 110 barg–160 barg. An optimum compressor discharge pressure exists at approximately 121 barg. Below this pressure, the increasing circulation rate dominates, increasing the power demand. Above this pressure, the increasing compression ratio becomes the dominant factor due to the diminishing returns in reducing the circulation rate. Note: The calculated compressor power is the hydraulic power (i.e., power at 100% adiabatic efficiency).
The question that follows is whether this optimum point is always at 121 barg or whether it is dependent on other variables, such as the chiller temperature or the air cooler discharge temperature.
FIGS. 4a–4f show plots of compressor discharge pressure against power for a range of air cooler temperatures (32°C–60°C), with fixed refrigerant temperatures ranging from 0°C–50°C. Red points indicate the minimum compressor power on each curve. These figures show the following correlations between chiller temperature, aftercooler temperature, compressor discharge pressure and compressor power:
There is also a correlation between the chiller’s temperature and the compressor’s power, although, in FIGS. 4a–4f, this relationship is not immediately apparent. FIG. 5 shows the optimum compressor discharge pressure curves for all chiller temperatures plotted against the aftercooler discharge temperature.
It is important to remember that the definition of optimum in this context is the pressure corresponding to the minimum power consumption of the compressor. As the discharge pressure of the compressor increases, limitations on compressor design may necessitate a multi-stage machine, and design pressure limitations may require changes in pipe class that can incur step changes in capital costs.
Above ~120 barg, there are only small reductions in compressor power for further increases in discharge pressure. The true optimum pressure should therefore be assessed on a case-by-case basis, considering the capital cost of the equipment. In a closed-loop system, it is unlikely that increasing the discharge pressure beyond ~120 barg will yield a capital cost reduction. However, in an open-loop system where the refrigeration loop is integrated with a multi-stage injection compressor, higher pressure stages may be available, in which case FIG. 5 provides a useful guideline as to which is the most efficient stage to remove the refrigerant.
Alternative refrigerants. Refrigerant selection is often constrained by the temperature required in the chiller, which is limited to the atmospheric boiling point of the refrigerant, given that it is undesirable to operate the suction of the refrigerant compressor at vacuum conditions. It is often preferrable to select a refrigerant with the highest boiling point that can still meet the chilling requirement, as this will enable the system to operate at a lower pressure. TABLE 2 shows the atmospheric boiling points of some commonly used industrial refrigerants.
FIG. 6 shows a pressure-temperature diagram with boiling point curves for CO2 and for the refrigerants listed in TABLE 2. Each curve terminates at the component’s critical point. For any temperature, the saturation pressure of CO2 is higher than that of all the alternative refrigerants shown. Given that the refrigerant will leave the chiller as a saturated vapor, this means that a CO2 refrigeration system will always operate at a higher pressure than any of these alternative refrigerants at the same temperature.
The boiling curve of CO2 does not extend down to atmospheric pressure (0 barg). Instead, it reaches its triple point at ~4.2 barg and –56.6°C. Below this point, the equilibrium is solid-vapor (sublimation) rather than liquid-vapor. If the J-T valve were to decrease CO2 to below this pressure, solid CO2 (i.e., dry ice) would form in the J-T valve and downstream pipework. Therefore, the limitation on the minimum achievable temperature for CO2 refrigerant is not the atmospheric boiling point (which, for CO2, does not exist), but the triple point.
For comparison, the triple points of the alternative refrigerants are shown in TABLE 3. All the alternative refrigerants listed have triple points that occur at deep vacuum; therefore, the formation of solid-phase refrigerant is of no concern.
Considering that some margin to the triple point would be required to mitigate the risk of forming dry ice in the system, CO2 could practically be used as a refrigerant for services down to around –50°C. This is comparable to the minimum temperature achievable using propane, R-134a, R-125 and R-32. Of these, propane is by far the most common refrigerant used in large energy and oil and gas projects, and as a result, can be simulated most reliably in common process simulation software. There is also increasing pressure to phase out HFC refrigerants due to their high GWP. Propane is a likely replacement for many of these HFC refrigerant systems given its low GWP and low boiling point. Therefore, propane will be the refrigerant to which CO2 will be compared in the following sections.
System depressuring and material selection. When a system contains a liquid that has a vapor pressure greater than atmospheric pressure, the potential for low temperatures to be generated by auto-refrigeration during system depressuring must be considered. Refrigeration systems are particularly prone to this phenomenon, as they often contain pure or nearly pure component liquids that are at their boiling points at elevated pressure.
In a pure component system, a singular defined boiling temperature corresponds to every pressure. If the system pressure falls below the fluid’s vapor pressure at the system temperature, this will result in vapor generation. The latent heat required to generate this vapor is lost as sensible heat, resulting in the system becoming colder. Vapor will be generated until the system pressure is equal to the vapor pressure at the new, lower temperature.
If the pressure continues to reduce due to mass being lost from the system—as is the case during system depressuring—this phenomenon will continue. As the pressure drops and liquid boils off, it will remain on its boiling point curve until it either reaches its atmospheric boiling point or until all liquid has been boiled off and all that remains is saturated vapor. If the former is the case, the system will remain with an inventory of liquid at its atmospheric boiling point until heat input from the environment can boil off all remaining liquid. Therefore, it is good practice to set the minimum design temperature of the system to be no greater than the atmospheric boiling point of the refrigerant.
For propane, the atmospheric boiling point is –42°C. This is fortuitously a small margin above the minimum design temperature that is typically specified for low-temperature carbon steel (LTCS), which is –46°C.12 As previously discussed, CO2 does not have an atmospheric boiling point; rather, it has an atmospheric sublimation point—however, the principle is the same as described above.
This means that during rapid depressuring, it is highly likely that solid-phase CO2 (dry ice) will form in the system at the atmospheric sublimation temperature of CO2 (–78.5°C). This design temperature cannot be accommodated by LTCS; rather, it requires alternative metallurgy suited to such low temperatures, such as stainless steel. Therefore, all equipment in a CO2 refrigeration system would require a step change in capital costs compared to an equivalent propane refrigeration system due to the more expensive material selection required.
The requirement to set the minimum design temperature to the atmospheric boiling (or sublimation) point can be alleviated if it can be justifiably argued that the system will not reach this temperature in practice. Consider that, as the system cools down, the temperature will fall below that of the ambient air, and it will begin to gain heat from the surrounding environment. If the system is depressurized sufficiently slowly, an equilibrium can be reached where the heat gain from the environment is sufficient to replace the heat lost in vaporizing the refrigerant. The slower the rate of depressurization, the higher the temperature will be at which this equilibrium can occur.
In hydrocarbon systems, there is often a code-imposed requirement to depressurize a system to a certain pressure within a defined timeframe. However, this requirement may not apply to CO2 systems as the fluid is non-flammable. It may be possible to depressurize the system sufficiently slowly so that the system never drops below the design temperature of LTCS; however, this must be proven on a case-by-case basis through rigorous dynamic simulation.
CASE STUDY: A CARBON CAPTURE PLANT WITH INTEGRATED CO2 REFRIGERATION
A CO2 refrigeration system would perhaps be best integrated with a carbon capture and injection plant where there is a significant refrigeration duty required. One such system would be a propylene carbonate carbon capture plant. Propylene carbonate is a physical solvent that is regenerated through multi-stage flashes with no heat input. These plants have no process heat demand but can have a significant refrigeration duty associated with chilling the lean solvent for improved rich-solvent loading in the absorber.13 The chilled solvent temperature is typically in the range of –25°C to –30°C with a very tight approach in the chiller. For this case study, a refrigerant temperature of –30°C was selected with a solvent temperature of –29°C.
A case study was performed, comparing a conventional propane refrigeration system to a CO2 refrigeration system to provide chilling duty to a propylene carbonate CO2 capture plant. The plant will capture CO2 on the basis shown in TABLE 4 and compress the captured CO2 for sequestration.
Two configurations for the CO2 refrigeration system were considered. The first was a closed-loop CO2 system as described in earlier sections. The second one was an open-loop system (FIG. 7) in which the refrigeration system was integrated with the CO2 injection compressor.
The CO2 injection compressor considered was based on a seven-stage machine that would increase pressure from 0.85 barg to 180 barg for injection, with intercooling to 45°C between each stage. The interstage pressures are shown in TABLE 5 and were based on a typical pressure profile.
Propane refrigerant. The vapor pressure of propane at 45°C is 14.4 barg. Allowing a 0.5-bar pressure drop for the condenser provides a compressor discharge pressure of 14.9 barg. The suction pressure is the vapor pressure of propane at –30°C, which equals 0.66 barg.
Closed-loop CO2 refrigerant. Using FIG. 5, the optimum compressor discharge pressure with a –30°C chiller and a 45°C air cooler is 142 barg; however, this would result in a high compression ratio and a high discharge temperature. From FIG. 4d, the power curve is quite flat beyond ~120 barg. In fact, the compressor power at 120 barg is less than 3% higher than that at 142 barg. Therefore, 120 barg was selected as the compressor discharge pressure, as this results in a discharge temperature of 170°C, which is high but within the range of a centrifugal compressor. The suction pressure is the vapor pressure of CO2 at –30°C, which equals 13.3 barg.
Open-loop CO2 refrigerant. In the open-loop configuration, the refrigeration loop was positioned in a recycle loop around several stages of the CO2 injection compressor. This has the benefit of eliminating the separate refrigeration compressor and aftercooler by incorporating these duties into the injection compressor and interstage coolers.
The disadvantage of this configuration is that the interstage pressures are unlikely to be optimum for running the refrigeration loop. This means that the additional compressor power will be greater than for the independent closed-loop refrigeration unit, which can be optimized for chiller requirements.
The closest pressure to the optimum of 142 barg is the discharge of Stage 6 at 120 barg; therefore, this will be the take-off location for the refrigerant supply (downstream of the interstage cooler). It is beneficial to return the refrigerant at the highest possible pressure that can meet the chiller temperature of –30°C (vapor pressure of 13.3 barg). The best option is to return the refrigerant to the suction of Stage 3 at 9.5 barg. TABLE 6 shows the results of simulations for each refrigeration configuration.
Takeaways. A propane refrigeration system operates at much lower pressures than an equivalent CO2 refrigeration system. The propane system can be accommodated by 300# piping (or even 150#) depending on the achievable condensing pressure. A CO2 system will likely require 900# or 1500# piping due to the high design pressure required. The issues with the depressurizing temperatures discussed mean that stainless-steel metallurgy will usually be required for the CO2 system, whereas the propane system can be manufactured from LTCS.
The equipment for the CO2 system will typically be physically smaller than the equivalent propane equipment. This is largely due to the higher operating pressure and reduced system volumes. It may also be possible to reduce the size and number of chillers required by using CO2, which can yield a higher heat transfer coefficient than propane.
To provide the same chilling duty, a CO2 system requires more power than a propane system. A closed loop is easier to optimize and can have lower power requirements than an open-loop system. If an open-loop system is used, it is highly beneficial to set the interstage pressures such that stages exist close to the optimum refrigeration supply and return pressures.
On balance, it is unlikely that a closed-loop CO2 refrigeration system will provide a cost-competitive solution vs. a conventional propane refrigeration system. However, if an open-loop system is possible, in which the refrigeration loop is integrated with a multi-stage CO2 injection or export compressor, this may present a lower capital cost solution than a standalone propane refrigeration system. This is especially true if the additional compressor duty is a relatively small percentage of the total injection compressor duty.
In addition, a CO2 refrigeration system has the following benefits:
ACKNOWLEDGMENTS
The authors would like to thank Adish Jain, Process Technology Director, Fluor, and Samantha Nicholson, Process Fellow, Fluor, for their valuable input to this article.
LITERATURE CITED
CHARLIE GOULD is a Process Engineer at Fluor. He has 8 yr of experience on a variety of upstream oil and gas, refining, chemicals, heat and power, carbon capture and hydrogen projects. He specializes in steady-state and dynamic process simulation, overpressure protection and process safety time analysis. Gould earned an MEng degree in chemical engineering from the University of Surrey in the UK and is a chartered engineer with IChemE.
ROBERT JOYNER is a Principal Process Engineer at Fluor. He has more than 20 yr of experience in onshore and offshore oil and gas, refining and chemicals projects, specializing in overpressure protection and flare system design. He earned an MEng degree in chemical engineering from Imperial College London and is a chartered engineer with IChemE.
ADYA DESHMUKH is a Senior Process Engineer at Fluor. She has more than 20 yr of experience, specializing in CO2 processing, including Econamine FG PlusSM, direct air capture and CO2 compression/conditioning process design. She graduated from the University of Surrey in chemical engineering and is a registered chartered engineer.