Design Considerations when Using Carbon Dioxide in Industrial Refrigeration Systems
Angus Gillies B. Eng, C. Eng &
David Blackhurst BSc(Hons), C.Eng
Star Refrigeration Ltd
Glasgow, UK
Abstract
R744 is a natural, efficient, safe and environmentally friendly alternative to more common refrigerants
for industrial refrigeration systems but problems in the detail of the design have prevented common
use. High pressures have limited the availability of suitable system components, particularly to allow
hot gas defrosting. Cascade operation overcomes many of the problems but it can still be a significant
challenge to find suitable components and an economic system to defrost a cascade refrigeration plant.
This paper considers the practical aspects of design and component selection when applying carbon
dioxide to large industrial cascade and volatile secondary refrigeration plant with hot gas defrosting.
Introduction
Concept aspects related to the use of carbon dioxide in industrial refrigeration plant are already well
covered, including concept design (Pearson, 2003) and selection for efficiency (Blackhurst, 2003).
This paper deals with the practical aspects of the design of cascade and volatile secondary carbon
dioxide refrigeration plant. Many of the challenges in adopting carbon dioxide as a refrigerant relate to
the relatively high pressures. Thus component selection and detailed system design have posed
significant challenges for those adopting carbon dioxide as a primary refrigerant. Adapting present
refrigeration design and introducing components from the petrochemical field have made the
introduction of cost effective carbon dioxide systems a reality.
This paper describes the lessons learned planning, designing and installing four cascade carbon dioxide
systems, two of which included volatile secondary circuits, average duty 2.6 MW. This paper assumes
R717 high stage plant. Current and future availability of components is considered to allow system
designers to best assess the options available.
System Allowable Pressures, Component Design and Material Requirements
Detailed consideration has shown that the economic design and easily achievable limits with currently
available components has dictated a maximum allowable pressure (EN378, 2000) of:
- 40 Bar(G) for all parts of a cascade circuit not exposed to defrost pressure when carbon dioxide
condensing pressures are below –4.0°C to allow a 20% safety margin (IOR Code, 2003(1)) and for
some margin for pressure fluctuation.
- 51 Bar(G) for all parts of a cascade system exposed to hot gas defrost pressures allowing a 10%
safety margin (IOR Code, 2003(2))
- 90 Bar(G) for carbon dioxide charging connections
The 40Bar(G) limit is derived from common component designs and allows use of standard or
modified industrial refrigeration components, pressures lower than this are un-necessary and are likely
to lead to increased incidence of high pressure cutouts tripping and refrigerant loss due to relief valves
opening on power failure. It is recommended that the section of the system normally only exposed to
the evaporation pressure is designed for the same pressure as the condensing section as its pressure will
need to rise to this value if the carbon dioxide compressors are not available to operate. We have good
experience with defrosting carbon dioxide tube and fin air coolers and vertical plate freezers at a
condensing temperature of 10°C.
Pipe and vessel material requirements will commonly mean selection of LT 50 (ASTM A333 GR6)
carbon steel, or equivalent, or stainless steel which may be an economic choice (Pearson, 1993). It
should also be noted that the common supply requirements of bulk carbon dioxide distributors will
require a vessel design temperature of -50°C even if the operating temperature is well above this value
(BCGA, 1999). Although de-pressurisation or low temperature operation of the system would cause
temperatures below -50°C this is acceptable with reference to many European design codes as the low
temperatures are co-incident with low pressures and therefore low membrane stress (PD5500:2003).
Refrigerant stop valves are already available, in Europe, with a design pressure of 52 Bar(G).
Tests were conducted with an American valve manufacturer for operation of modified versions of
standard refrigeration solenoid and pressure regulating valves with carbon dioxide at up to 40 Bar(G)
and this has been successful in normal operation. However, we have found in testing that solenoid
valve diaphragms have been damaged during testing to simulate repeated operation at differential
pressures of approximately 40 Bar.
Ball valves used in process industry are suitable for use with carbon dioxide as a refrigerant although
they do not have glands that can be changed under pressure. These are commonly available as a 3
piece design with design pressures of PN63 and sizes up to DN80 for full bore models. We
hydraulically tested a sample to 255 Bar(G), 5 times our system allowable pressure, without problem.
Glass reinforced PTFE seats have been found to work well. Graphite-filled PTFE seats tend to require
excessive torque to operate the valve even at low differential pressures; this may be related to the low
moisture content of the carbon dioxide used in refrigeration plant. Ensure all ball valves are of the
vented cavity design to prevent high pressure build-up inside the cavity of a closed valve.
Refrigerant charging connections should be designed for 90 Bar(G) up to and including the first stop
valve because of the danger of over pressurisation if the valve on the system is closed when charging
from a carbon dioxide cylinder at ambient temperature. Ball valves up to DN 25 are easily available
for this duty.
It is considered of primary importance that components especially adapted to higher pressures by
changing materials and test regimes but not basic design are very clearly marked to highlight that they
are “specials” to avoid future mix-ups. This should be carried out before the valves are delivered to
the manufacturing facility, for example by having the valve bodies painted a non-standard colour.
Cascade Condenser and Associated System Design Considerations
Consideration of realistic cascade heat exchanger sizing is likely to start with heat exchangers sized for
2.5°C difference however as wide as 6°C can be economic if the payback period required is short or
mean operating load is well below design.
The potential risk of high pressure refrigerant in the low temperature stage of a cascade system
entering the high temperature stage through a leak in the cascade condenser must always be assessed
by the designer. Special consideration of the pressure build-up in an ammonia plant due to carbon
dioxide leaking into it are necessary because the two fluids will mix to form ammonium carbamate
which is a solid with potential to cause great harm including blocking relief valves. The author has
demonstrated this in a controlled test in a pressure vessel. Clearly the selection of this heat exchanger
should consider likelihood of leakage as a major factor.
Two methods of reducing the risk of mixing working fluids have been used. Firstly motorised valves
can be used to shut off the connections of the cascade heat exchanger in the event of an indication of a
leak. This can be triggered by an ammonium carbamate detector, such as an optical device detecting
the presence of the white powder, or pressure switch sensing the evaporating pressure in the ammonia
system. The former option is preferred as the pressure in the ammonia circuit will initially fall as the
production of ammonium carbamate consumes both ammonia and carbon dioxide. Incorporation of
both sensing systems is a recommended. Valves on either the carbon dioxide connections, ammonia
connections or both would be appropriate but detailed assessment of both harm and costs leads the
author to suggest automatic valves on the carbon dioxide connections only for systems with a common
carbon dioxide circuit and twin ammonia circuits. Secondly increased standard refrigeration
component design pressures combined with design carbon dioxide condensing pressure below
approximately -15°C will allow economic construction of the complete plant to a single maximum
allowable pressure ensuring that the operating pressure of the carbon dioxide plant will not cause overpressurisation of the ammonia plant due to leakage in the cascade heat exchanger and blockage of the
ammonia plant’s relief valves.
Economic cascade heat exchanger selection is likely to include the choice between shell and tube, plate
and shell and fully welded plate and frame heat exchangers. Although shell and tube designs are
widely available and can be built with a double tube sheet design to limit the possibility of cross
contamination between refrigerants current designs are relatively large and heavy with high ammonia
charges. Experimental tests and experience also shows that heat transfer coefficients are relatively low
at low duty so system efficiency benefits at low duty are not as may be expected. There is some
evidence that this is due to slow draining of condensed carbon dioxide inside the tubes so inclined tube
designs should be considered. Plate and shell and fully welded plate and frame heat exchangers
largely overcome the disadvantages of shell and tube designs with the choice between the two being
governed by the likelihood of leakage and local commercial considerations.
Concern over the potential failure of a cascade heat exchanger has led to two of the systems built so far
to include two separate ammonia circuits. It seems likely that this design will become more common
as standard packaged carbon dioxide condensing units could be marketed in the same way as standard
packaged water chillers. This offers the following advantages with minimal extra cost or reduction in
efficiency:
- Leakage of carbon dioxide to ammonia due to a leak in one cascade heat exchanger would not
mean loss of all refrigeration capacity
- There is no need for an auxiliary plant to hold carbon dioxide pressures should one ammonia
system fail to operate
Defrost Systems
Cascade refrigeration plant will typically be designed with carbon dioxide condensing pressures in the
range -20°C to -5°C so hot gas defrost using the compressor discharge gas is not possible. Currently it
seems elevating the carbon dioxide condensing pressure to that possible for hot gas defrosting will be
difficult due to component limitations and even then it may not give best plant efficiency. Two options
to generate gas pressures suitable for defrosting are currently in use. Firstly a compressor can be used
to generate the gas pressure (Nielsen, 2003) and although this is a relatively straightforward option
with good system efficiency no refrigeration compressor, available on the open market, is capable of
such pressures. Even if these become available they are likely to be at the extreme of their operational
limits and require high maintenance. Petrochemical type screw compressors with cast steel bodies are
suitable for the application but the costs and delivery times are outside the normal limits for industrial
refrigeration plant.
We have built three successful systems using high pressure pumps to boost carbon dioxide pressure
from that at the condenser outlet (Figure 1) to the pressure required for defrosting then evaporating and
superheating the carbon dioxide using waste heat from either the glycol oil cooling circuit or
condensing ammonia. This is being used for defrosting air coolers and plate freezers. System
advantages include:
- Defrost pressures are only limited by valve, heat exchanger and pipe design pressure limits.
- The system is easily controlled
- Availability of appropriate pumps proven in carbon dioxide applications is good
- Pump maintenance is low and reliability is excellent
- Pumps can operate well within their design limits
- Operation near the cutout pressure is stable with good reliability
The use of the glycol is a good proposition as long as it can supply sufficient waste heat as it
significantly reduces the chance of cross leakage between the carbon dioxide and ammonia; which is
more likely on a defrost vapour generator than a normal cascade heat exchanger because of the
additional stresses from thermal and pressure cycling.
Figure 1, Defrost Vapour Generator
The pumps used are single stage positive displacement plunger type. It should be noted that these
pumps require sub-cooled liquid at inlet to prevent liquid flashing to vapour due to pressure drop in the
inlet valve, this has been successfully achieved by feeding them from centrifugal pumps or mounting
the pumps approximately 6m below the storage vessel or feeding them with a sub-cooler. These
pumps have not been tested at fluid entering temperatures below -30°C where formation of frost on
piston rods could damage shaft seals.
Pressure is controlled by spilling the pumped liquid from pump discharge to pump suction via a back
pressure control valve with mechanical pilot (valve A) thus tending to starve the defrost evaporator of
refrigerant and quickly reducing the rate of gas generation. Should the rate of usage of gas exceed the
waste heat available the level in the separation vessel will rise in which case we regulate liquid feed to
the evaporator. The superheater is included to prevent condensation and the risk of liquid hammer in
the hot gas line.
Air cooler circuit design for defrosting does not need any particularly special consideration, we have
had good experience with stainless steel tube aluminium fin coolers with 16 mm tube diameter and
rising circuits and hot gas drip tray heating with gas feed at 20°C temperature and 10°C saturated
pressure and with room temperatures down to -26°C. The tray design should give low pressure drop
and should include tubes near the edges so that sufficient heat is conducted to these points as the gas
temperature is relatively low compared to most hot gas systems.
The design used in these systems incorporated a defrost pressure release valve which is a modified
pressure relief valve. The seat is designed for pressure control, not pop action, and a bellows seal is
fitted.
Compressor Type
The choice of compressor for carbon dioxide applications is limited. Our best experience has been
with oil injected twin screw compressors that have been reliable with capacity and efficiency equal to
that predicted.
With any compressor choice it is likely that maximum suction pressure limits will be relatively close
(in °C) to the design operating pressure so some means of suction pressure regulation will be required
on many systems. Manual throttling of the suction pressure control valve has proven inadequate on
one carbon dioxide system.
Vessel Sizing
Vessel sizing based on ASHRAE droplet separation principles (Refrigeration Handbook, 1998) has
been successful and it will be recognised that relatively low gas volume flow rates mean smaller
vessels. However the gas density is high so separation velocities are low and liquid volumes are likely
to be equal to or even greater than an equivalent ammonia system due to large liquid line volumes and
high liquid fractions in the evaporators. In addition the mass of refrigerant in the gas phase cannot be
considered negligible as in some systems it could be over 20% of the total system charge.
Air Cooler Circuit Design
Performance testing has shown us that high pressure drop circuit design, exceeding original predicted
optimum values, gives good air cooler operation. The high vapour density means the liquid / gas
fraction by volume in the mixture coming off the coil is relatively high when compared with other
working fluids so the overfeed rate doesn’t need to be as high to achieve adequate wetting of the inner
tube surface. In fact even lower overfeed rates may be possible however low inlet velocity and
imperfect distribution between circuits will act against this. We started with a very conservative
pumped feed rate but have found good evaporator performance with much lower rates and suggest the
following:
Condition
Carbon Dioxide 0°C Air Cooler
Carbon Dioxide -10°C Air Cooler
Carbon Dioxide -20°C Air Cooler
Carbon Dioxide -30°C Air Cooler
Carbon Dioxide -40°C Air Cooler
R717 -40°C Air Cooler-comparison
Pumped Feed
Evaporation Rate
1.5
1.6
1.8
2.0
2.4
4.0
Rate/ Liquid Portion at Cooler Outlet
vol/vol (%)
5.0
4.2
3.9
3.3
3.2
0.28
Carbon Dioxide Pumps
R744 pump selection needs careful consideration, apparent sub-cooling due to gravity head above the
pumps is typically one fifth of that for ammonia at the same temperature thus increasing the chance of
flashing and therefore cavitation in the pump. In addition the surface tension of carbon dioxide can be
as low as only 20% of that for ammonia depending on the temperature so vapour bubbles form
differently and the relatively low viscosity means the lubricating effect is not so good. A pump
supplier who understands these peculiarities and has experience with carbon dioxide should be sought.
The system designer is advised to take a conservative approach when considering the pump suction
head required and pump suction line sizing.
Beware of pressure drop in the evaporator and the wet suction line as this could add 10m to the pump
head required. Further the relatively large change in saturation pressure per °C effectively precludes
use of back pressure control of individual evaporators on pumped overfeed carbon dioxide systems.
Piping Design
Vapour pipe and valve size savings can be considerable when using carbon dioxide (Pearson, 1993)
and considerable secondary savings can also be made when the designers recognize the further benefits
in insulation, pipe weight, installation time and space requirements. Wet return piping with carbon
dioxide needs some special consideration as the allowable pressure drop (in °C) is likely to be less than
for other refrigerants due to pump head limits (refer above) and cost effective design considerations.
Assessment of economic pipe sizing (Stoecker, 1998) will show that the allowable pressure drop (in
°C) for carbon dioxide systems will be lower than for other common refrigerants because the cost of
reducing the pressure drop is relatively small. Typical wet return pipe velocities will be lower than for
other refrigerants for the reasons above but also because the liquid fraction (by volume) is higher,
based on the overfeed rates proposed above.
The Process Associates website
(www.processassociates.com) is recommended as a useful horizontal wet return pipe sizing tool. It
should be noted that typical rules for wet return riser velocities are not applicable due to the high gas
density (velocities should be lower even though liquid fractions are higher) but that wet return risers at
low loads do not pose the same problems because the pressure loss (in °C) due to a liquid column is
only approximately 17% of that with R717; and the density ratio between vapour and liquid is low.
Low pressure drop (in °C) on plate freezer wet return flexible hoses is also another significant benefit
with carbon dioxide.
Oil Type
Both polyolester and polyalphaolefin are suitable for use with carbon dioxide systems. POE has the
advantage of miscibility with the refrigerant (Watson, 1998) making oil return much easier; however
the author has experience of an as yet un-explained screw compressor failure with this oil and carbon
dioxide. Use of this oil is likely to be common in the future once more experience is gained. PAO has
the disadvantage that it is not miscible with the refrigerant but the advantage that there is long and
good experience with this lubricant in carbon dioxide compressors in process plant (O’Neill, 1993).
The systems we have built recently have all used PAO oil as this is considered the best solution for
long term, reliable, compressor operation. It also allows the same oil to be used in the ammonia and
carbon dioxide compressors of a cascade plant. However as well as not being miscible with the
refrigerant it has the further disadvantage of being less dense than the liquid refrigerant at temperatures
below approximately 10°C so will tend to accumulate on the liquid surface of any low temperature
vessel. This has been countered by installing high efficiency oil separation with a second set of
coalescer elements with particle removal down to 0.01 micron to reduce the oil carry-over to a
minimum. Less than 2 ppm oil carry-over (by mass) is considered easily achievable with this system.
The liquid oil separator design proposed (Figure 2) allows collecting of oil which is buoyant in the
refrigerant although good oil separation means it is too early to comment on the effectiveness of this
device. Initially this has been installed in the liquid line from the condensers but as this only gives one
chance to catch the oil and evidence of oil has been found in the low temperature pumped liquid line it
may be better mounted there. Valves on the low temperature vessel to allow future skimming of the
liquid surface are considered a worthwhile inclusion.
Figure 2, Liquid Oil Separator
Control
System reliability for a large cascade plant with carbon dioxide significantly benefits from the
inclusion of discharge pressure capacity limitation of the carbon dioxide compressors. Relatively strict
limits are put on compressor operation and a modern control system should be used to limit evaporator
operation to keep within those limits. Control of high stage compressors based on carbon dioxide
condensing pressure has proven stable and efficient.
Relief Valves
Carbon dioxide can form solid in a relief valve even if there is only vapour at inlet (above
approximately18 Bar(G) valve set pressure) so relief valve mounting should be at the end of the high
pressure line from the vessel. The maximum pressure drop before the valve should be limited
according to current regulations (EN13136, 2001).
Safety
The author recommends that a full risk assessment is carried out and that a gas detection system is a
sensible precaution for pits and low areas where carbon dioxide could accumulate and plant rooms and
possibly roof voids where valve stations are mounted. Infra red type gas detection systems utilising a
silicon based sensor structure have shown to be robust and reliable without giving false alarms. We set
these to 5000ppm (the 8 hour TWA limit) to give an early warning of potential problems.
Refrigerant Quality
The standard food grade of commercially available carbon dioxide has been found to be acceptable for
refrigeration use. This has a moisture content of <10ppm by volume in liquid carbon dioxide.
Molecular sieve type driers should be incorporated to minimise the chance of corrosion or reaction
with POE oils.
References
1. Pearson, A.B., Cable, P.C. (2003), A Distribution Warehouse with CO2 as Refrigerant, The
Proceedings of the 21st IIR International Congress of Refrigeration, ICR064
2. Blackhurst, D. (2002), CO2 v NH3, A Comparison of Two Systems, The Proceedings of the
Institute of Refrigeration, 2002-03, 2-1
3. EN378 (2000), Part 1:Basic requirements, definitions, classification and selection criteria,
Refrigeration systems and heat pumps-Safety and environmental requirements, chapter 3.3.2, p7
4. IOR Code 2003(1) Safety code for refrigeration systems utilizing carbon dioxide, Appendix B,
Table B1.
5. IOR Code 2003(2) Safety code for refrigeration systems utilizing carbon dioxide, Appendix B, Note
6.
6. Pearson, S.F. (1993), Development of Improved Secondary Refrigerants, The Proceedings of the
Institute of Refrigeration, Volume 89, 1992-93, p.72-77
7. BCGA (1999) British Compressed Gasses Association Code of Practice CP 26 Bulk carbon
dioxide storage at users’ premises, Rev1, section 1.2.2, p.5-6
8. PD5500:2003, Specification for Unfired Fusion Welded Pressure Vessels, Annex D, Section
D.5.1.1
9. Nielsen, P.S., Lund, T. (2003), Introducing a New Ammonia / CO2 Concept for Large Fishing
Vessels, 2003 Ammonia Refrigeration Conference and Exhibition Technical Papers, IIAR,
Technical Paper #11, p.372-374
10. Stoecker, W. (1998), Optimum pipe size, Industrial Refrigeration Handbook, chapter 9.5, p.345346
11. Watson, M., Oberle, J., Rajewski, T. (1998), Lubricants for Natural Working Fluids, Applications
for Natural Refrigerants, Refrigeration Science and Technology Proceedings, p.407-415
12. EN13136 (2001), Pressure loss in upstream/downstream lines, Refrigerating systems and heat
pumps- Pressure relief devices and their associated piping- Methods for calculation, chapter 7.4,
p.10
13. O’Neill, P. (1993), Industrial compressor lubricants, Industrial Compressors, chapter 10.8.2, p.245