Co2_brine

Technical Papers34th Annual Meeting

International Institute of Ammonia Refrigeration

March 18–21, 2012

2012 Industrial Refrigeration Conference & ExhibitionHilton Milwaukee City Center

Milwaukee, Wisconsin

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ACKNOWLEDGEMENT

The success of the 34th Annual Meeting of the International Institute of Ammonia

Refrigeration is due to the quality of the technical papers in this volume and the labor of its

authors. IIAR expresses its deep appreciation to the authors, reviewers and editors for their

contributions to the ammonia refrigeration industry.

Board of Directors, International Institute of Ammonia Refrigeration

ABOUT THIS VOLUME

IIAR Technical Papers are subjected to rigorous technical peer review.

The views expressed in the papers in this volume are those of the authors, not the

International Institute of Ammonia Refrigeration. They are not official positions of the

Institute and are not officially endorsed.

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2012 Industrial Refrigeration Conference & Exhibition

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© IIAR 2012 1

Abstract

Calculations and analysis of practical case studies indicate that systems using CO2 as a secondary coolant provide energy savings in the range of 15 to 30 percent compared to their glycol, brine counterparts. Those systems are also relatively simple. The development of CO2 brine systems is rapidly accelerating, with end-users in a number of countries opting for those plants. This paper outlines one of the challenges faced by many contractors – the design differences of CO2 secondary cooling systems as compared with traditional ammonia installations, or glycol layouts.

Technical Paper #7

System Design for CO2 Secondary Coolant Based System

Hernan HidalgoDanfoss Inc.

Baltimore, Maryland

Anatolii MikhailovDanfoss A/S

Kolding, Denmark



Technical Paper #7 © IIAR 2012 3


Introduction

At the dawn of the 21st century, industrial ammonia refrigeration systems

experienced rapid changes while embracing other re-emergent natural refrigerants

such as CO2 which in turn, helped realize potential benefits to product safety and

operational effectiveness.

The evolution and variants of CO2 system design range from CO2 cascade, to DX

and brine systems or a combination. All of them have been used to fulfill the

requirements of low and medium temperature systems. This paper focuses on

CO2 secondary systems addressing the control challenges in some medium to high

temperature applications where the use of traditional ammonia or glycol based

systems is still widely common practice.

This paper provides the basics of CO2 secondary system designs. Without being a

complete guide, the article creates a general understanding of the challenges and

opportunities that application of CO2 as secondary coolant creates. Most of the

high and medium temperature applications that use glycol as a secondary cooling

today, could successfully apply CO2 instead, with benefits for the environment and

potentially the energy bill. The latter will be explored and reviewed later in the paper.

Traditional Systems Overview

Ammonia based systems are the preferred choice for industrial refrigeration system

designs. These systems provide a well known, predictable and easy-to-service layout

which has very desirable performance while keeping environmental standards. There

are, however, a significant number of applications where the use of ammonia/glycol

systems may be preferred.


4 © IIAR 2012 Technical Paper #7

2012 IIAR Industrial Refrigeration Conference & Exhibition, Milwaukee, Wisconsin

NH3 Direct

While most systems operate at +20°F (–6.7 °C) evaporating temperatures at the

pump re-circulator package, high temperature units require the use of pressure

regulators to increase evaporating temperatures to achieve higher room temperatures

with an acceptable TD (temperature difference).

Though the practice is not uncommon, it brings some disadvantages in terms of

pump power required to overcome the setting of the pressure regulator as well as

energy consumption by the compressors.

Figure 1 shows a pumped ammonia valve arrangement for medium and/or high

temperature units. The pressure regulator shown in the wet return line is set at the

desired higher evaporating temperature. A unit operating at 45°F would require an

evaporating temperature of 35°F. The setting of the pressure regulator will around

55 psig. As +20°F saturated liquid would have a pressure of ~ 34 psig, the ammonia

pump (s) require to have a differential pressure in excess of 30 psig just to feed liquid

to those units. Due to the high differential pressure required some designers may

have opted for high pressure liquid instead of medium temperature pumped liquid.

The down side beyond having higher power consumption at the pumps is the

recirculation of ammonia in process areas (with potentially a high number of

employees) or sensitive products susceptible to damage even with very low

concentrations of refrigerant in the room.

NH3/glycol based

Those hybrid NH3/glycol systems aim at addressing risk management of plants where

product contamination and/or extensive process areas where companies with very

low tolerance for potential ammonia leak exposure operate. The ultimate goal is to

reduce exposure to leaks with NH3 which, of course, is highly desirable.




NH3/glycol systems (or any other refrigerant with glycol) accomplish the task.

Nevertheless, there are substantial repercussions plant managers have to contend

with as the systems:

• Requirealargefootprint.Glycolpipelinesareratherlarge,especiallywhen

insulationistakenintoaccount.Glycolsystemsalsorequireratherlargevessels

(or dual vessels) to separate warm and cool liquid.

• Pumppowerisextremelyhighwhichispresentonenergybillsmonthafter

month.

• NeedtomonitorandmaintainGlycolmixtureproperties.

Figure 2 shows a hybrid refrigerant/glycol system.

Warm glycol is pumped through a chiller (NH3/Glycolcommonlyused).Itisalso

common to have another pump (system pump) feeding the cooler units in the plant.

The arrangement increases power consumption and complexity of the plant.

Motorized three way mixing valves with a small re-circulating pump are typically

used to obtain the right glycol temperature. This strategy is used when precise

temperature and / or humidity control is required. In other cases, a simple on/off

control method is quite common.

CO2 as Brine Overview

Traditionally, CO2 systems were built for low temperature applications due to the

high efficiency associated with it. However, plants that have a demand for both low

and medium temperature requirements could also benefit from applying CO2 to the

medium temperature loop such as process areas, warehouses, banana rooms, etc.




The medium temperature side is typically a pumped circulation loop connecting

the cascade heat exchanger to the rest of the system. Nevertheless, stand alone

installations with a CO2 medium temperature loop are also viable.

Figure 3 illustrates a conventional CO2 system showing the medium temperature loop.

The vessel underneath the heat exchanger serves as a liquid/gas phase separator. The

returned vapor flows up to the exchanger by thermo-siphon effect to the exchanger.

This design makes control of CO2 brine systems much easier than those used for CO2

cascade systems:

• ThereisnoneedtobalancetheloadoftheCO2 and NH3 compressors in the

cascade heat exchanger. In CO2 brine systems the balance occurs automatically.

• Thesystemcouldoperateoil-free,whicheliminatestheneedofelaborateoil-

return systems.

However, even though CO2 brine systems do not operate with any compressors, they

feature a lot of similarities to traditional cascade systems.

Figure 4 illustrates a simplified CO2 system and the correlation of the different

conditionsoftherefrigerantinthelogP-Hdiagram.

The CO2 pump raises the pressure with a slight increase in the refrigerant temperature

due to mechanical losses (1-2). Further friction losses and head pressure positions

CO2 on point (3).

When the refrigerant reaches the high pressure rated pumped liquid valve station

containing shut off valve, strainer, pulse width modulating valve and shut off valve

modules, the refrigerant reaches the saturated liquid state. (4)

The refrigerant then evaporates and depending on the recirculation rate, wet vapor

returns to the CO2 receiver to complete the loop. (6)




CO2liquidandgasphasesareseparatedinthereceiver.Gas(7)flowstotheheat

exchanger where it condenses returning as slightly sub-cooled liquid. (8)

Liquid in the vessel remains in its saturated state. Some degree of sub-cooling can be

achieved depending on the liquid column height at the pump suction. (1)

Energy Efficiency of CO2 brine systems

Certainly, indirect ammonia systems have dramatically reduced the ammonia charge

by the use of secondary coolants such as CO2 volatile brine and glycol. However,

CO2 brine offers additional benefits such as increased efficiency compared to glycol

systems.

One key advantage of NH3/CO2 heat exchangers compared to other fluids is the very

high heat transfer coefficient available. (Handschuh, 2008) This applies to both air

units and heat exchangers.

The gains in evaporator performance are largely dependent on the dimensioning and

circuiting of the evaporator. Some comparisons of carbon dioxide coil design show

some variations of the heat transfer coefficient. A CO2 coil design with 10 circuits

instead of 20 circuits using the same overfeed rate could raise the heat transfer

coefficientabout9percent.(Pearson,IIAR2009)

It should also be mentioned that the lowest heat transfer in air cooled evaporators

is on the air side. Therefore, the attention will be placed on the refrigerants heat

exchange. As the temperature difference (TD) between CO2 and the other fluid (e.g.

ammonia) is quite small, the overall gains in operating higher compressor suction

(NH3 side) may range from 1.8°F to 7.2°F (0.5°K to 4°K). (Stenhede, 2007) As a

result, the energy consumption by the NH3 compressors would drop.




Furthermore, properties such as high specific heat and high vapor density of CO2

make an ideal combination to reduce line sizes and optimize components throughout

the refrigeration system design. Comparison of line sizes between CO2 and other

refrigerants has been done extensively proving large differences worth highlighting

again. (Vestergaard 2004)

Figure 5 shows dramatic reduction in pipe sizes for both supply and return of

medium and low temperature systems. Smaller CO2 pipes mean lower surface areas

which reduce the heat loss compared to larger brine/glycol pipes. (Hinde, 2009)

Another advantage is the lower pumping power required which will be discussed in

more detail.

Energy Consumption

Traditional system designs call for recirculation rates of 3:1 to 4:1 with NH3. The

required recirculation rate for CO2 to obtain the same or higher heat transfer

coefficientislower.Re-circulationratesvaryfrom1.1to2only.(Handschuh,2008)

The vastly lower recirculation rates reduce the energy to circulate CO2 through the air

units. However, the main contributor to having a low pumping power is in fact that

CO2 is volatile brine.

For example, when using other fluids such as glycol, the temperature of the fluid

as it passes the cooling unit will be raised between 10°F to 15°F (5°K to 8 °K). As

a result, a significant mass flow circulating the unit is required to achieve the same

effect compared to using CO2.

CO2 pumps, on average, consume only 5 percent to 10 percent of the energy required

to pump water based brines while achieving the same capacity. This also holds




true even though CO2 systems operate higher head pressures than glycol systems.

(Mikhailov, 2010)

The following analysis attempts to provide an overview of the potential savings.

Mass flow Comparison

A relatively simple calculation provides the mass flow of CO2 and the mass flow of

ethylene glycol required to transfer a certain amount of energy.

For ethylene glycol, the following equation is used:

M = Q / (C ∆T)

Assuming:

Glycoltemperature:–10°C,

C=3.42

∆T=4 °K, which is typical for this kind of systems

Because CO2 evaporates, a different formula is used:

M =Q/∆h,

The latent heat for CO2 at –10°C: 260 kJ/kg

Now, calculating the mass flow needed to reject 260 kW of cooling load under the

above operating conditions:

CO2: 260kW/260kJ/Kg ~ 1 kg/s

Ethylene glycol: 260/(3.42x4) ~ 20kg/s




Only one kg/s of CO2 vs. 20 kg/s of glycol is required.

This highlights that the mass flow of CO2 required to obtain the same cooling effect

becomes only a fraction of the mass flow used by traditional brine fluids.

Pump Power Comparison

It has been shown that the mass flow required to obtain certain cooling effect is

substantiallylargerwhenusingEthyleneGlycolcomparedtoCO2. The next step

would be to estimate the pumping power required.

The following example illustrates the difference in pumping power for a hypothetical

plant.

Assumptions:

PlantCapacity:Qo=500kW(142TR)

CO2 re-circulation rate: 1.5: 1

Differential Head: 25 m (82 ft)

∆T = glycol temperature difference, inlet/outlet = 4°K

ηco2 = 75%

A basic equation to calculate power consumption at the pump:

Ppump = q ρ g h / (cp 106)/ η

Ppump = (h Q0 9.81) / (∆T cp) / η




Pumpefficiencyisanempiricequationandcalculatedforaspecificpumptype.Itisa

function of the volumetric flow and the temperature:

η = 12,333*ln(V) + 14,144

V = ρ 3600 / m

m =mass flow, kg / s

m = (Q0 / cp)/ (tout – tin)

Ps CO2 = (h Q0 9.81 n) / r / η

The calculations show the power required to pump CO2 could be about 15 times

lower than the power used in pumping Ethylene glycol.

The results of pumping power required for various brines are shown on Table 1

under the same conditions. Two different temperature scenarios were considered,

–10°C (+14°F) and –20°C ( –4°F )

It is interesting to see how the power consumption increases for all other brines as

the temperature decreases.

Nomenclature:

M: [mass flow in kg/h]

Q: [Load of the cooler unit. kW ]

C: [Specific Heat of glycol in kJ/(kg*K). ]

∆T: [Temperature difference between glycol inlet and outlet in °K]

∆h: [Enthalpy of vaporization ]

Ppump = pump shaft power (kW)




q = flow capacity (m3/h). Calculated

ρ = density of fluid (kg/m3). Depends on the glycol

g = gravity (9.81 m/s2)

h = differential head (m).

η = pump efficiency. Estimated

cp=GlycolSpecificHeat,kJ/(kg*K).Dependsonglycol

V = volumetric flow, m3/h

r = heat of evaporation, CO2, depends on the temperature and pressure

Total Energy Consumption

A simplified assessment of the energy consumption of a refrigeration plant will be

used to compare CO2 vs. glycol systems. It is deemed simplified as some variables

such as energy required for defrost sequences etc. are not included.

Instead, the total power would be based on compressors and pump consumption.

Prequired=Pcompr+Ppumps

Ppumps Calculation procedure as shown on previous section.

Pcompr = Q0 adjusted/COP

COPofthecompressordependsonthetemperaturelevel.Forcalculationpurposes,

itwasassumedthattheCOPwouldbeequalto1.9for–20°Cairand3.8for0°Cair.

Compressor data was taken from calculation software of one of the major industrial

refrigeration manufacturers.

Q0 adjusted = Q0 + Q additional heat gains + Qpump




Q additional heat gains = radiation heat gains, 5% for CO2 and 7% for glycols due to larger

pipe dimensions.

Q pump = heat from pumps.

Another example of a 750kW plant will be reviewed with room air temperature of

5°C (39°F)

Table 2 shows the variables and assumptions made to calculate the total daily energy

consumption. A hefty 21 percent energy savings figure provides a promising result to

consider CO2 as an alternative to glycol.

Simulationtools(PackCalculationIIandCO2 Brine Calculator 3.91) have been

crucial to compare CO2 and brine systems performance. Following various results, it

can be argued that depending on the system configuration, CO2 brine systems run at

15 percent to 25 percent lower energy consumption than water based systems in a

temperature range of –40°F to +32°F (–40°C to 0°C).

Those figures are confirmed by some other studies (Natural Working Fluids in

ArtificialSkatingRinks,E.Berends).

The actual percentage depends on a number of variables such as:

• CO2 brine systems offer a lower reaction time making adjustments to suction

temperature when needed. Instead glycol systems require longer times to obtain

desired temperatures due to the high inertia of the system

• Loadpattern

• TypeofDefrost

• ControlStrategy

• Useofvariablefrequencyconverters(VFDs)

• Geometryanddimensionsofevaporators/coolers




Actual Power Consumption Comparison

In order to validate the theoretical calculations, data extracted from two refrigeration

plants were analyzed. It is worth mentioning that the comparison of energy

consumption of industrial refrigeration plants presents some challenges as the exact

cooling load is typically unknown and needs to be estimated. Two sites running on

similar load profiles and temperature levels were chosen to conduct the analysis.

The approach taken to analyze the data was made by obtaining the gross energy

consumption measured at distribution centers for fruits located in two different

climates in Europe. Both plants run ammonia as the primary refrigerant equipped

with evaporative condensers and floating head pressure control. As one plant is

running glycol and the other CO2 as brine, it can be contended that the comparison

offers valuable findings and insights. Table 3 provides a side to side comparison for

both plants.

Some calculation adjustments had to be made to offset the climate influence.

Otherwise, the comparison would favor the system located in the North.

The total consumed power in kW h is a measured figure that corresponds to 1.3mW

h for the glycol plant and 2.7mW h for the CO2. Though the values are different, the

purpose is to compare the consumption in kW / kW h.

The total difference in per unit consumption amounts to about 32 percent. That is

quite substantial but a margin of error should be taken due to assumptions in power

consumption of fans, lights, defrost strategy etc. can account for a deviation. Both

systems could improve energy efficiency by following several steps such as use of

frequency converters.




It must be noted that absolute conclusions of energy consumption cannot be derived

from energy data taken from only two sites as a number of assumptions have been

made compared to theoretical values.

CO2 Brine System Control for High Temperature Rooms.

It was highlighted that handling high temperature rooms with traditional systems

such as NH3 and pumped glycol bring some inefficiencies. NH3 re-circulated systems

would rely on standard pressure regulators to increase the refrigerant pressure and

temperature increasing pumping power.

Glycolsystemswith3wayvalves,typicallyusedinfruitstorage,providegood

temperature control but they are somewhat more complex, increasing the cost

compared to other available alternatives.

In the case of CO2 pumped systems, the penalty for increasing refrigerant pressure to

reduce the TD of the coolers/evaporators is simply not an option. Having a regular

on/off control strategy works well for all medium temperature units but poses a

challenge to products that require a lower TD to reduce weight loss etc.

Fortunately, the advancement in flow control and algorithm control strategy makes

it possible to solve the issue avoiding power consumption penalties while keeping

system pressure within acceptable levels.

On/off Flow Control individual components

In traditional pump circulation systems, the liquid feed is controlled by a thermostat

oranalogsignaltoaPLCwhichconstantlymeasurestheairtemperature.The

solenoid valve is opened for a period of time (several minutes) until the air




temperature reaches the desired set point. The mass of the refrigerant flow is

constant while the liquid feed process takes place.

This method offers a very simple way to control air temperature. However, large

temperature fluctuations caused by pre-defined differentials may cause undesired

effects in some applications where dehumidification and accurate temperature control

cannot be maintained.

Typical liquid valve groups include a solenoid valve and a hand regulating valve

(known as HEV in ammonia systems) which provide a fixed amount of refrigerant

which cannot be changed remotely.

Valve stations with pulse width modulation technology in pump re-circulated coolers

Adjusting the liquid feed to the actual capacity needed offers multiple benefits. The

air temperature is constantly measured and compared to the set point. When the

desired temperature is reached, a 754 psig pressure rated valve station capable of

pulse width modulation reduces the “opening” degree by reducing the time of liquid

injection based on a fixed cycle. See Fig 6.

The liquid refrigerant circulation rate is reduced when a lower amount of liquid

refrigerant is needed. As a result, a certain amount of “super heated gas” is obtained

in the air cooler, which in turn, reduces the ∆T between the refrigerant and the air.

Typical applications include fruit and vegetable storage where adjusting liquid feed is

required to match the actual load.

Figure 7 shows a graph of the temperature variations when using pulse modulation

technology.

In some large coolers, the use of motorized valves with a pressure rating of 754 psig

(52 bar) is also widely used.




Supporting Systems Overview

While major discussions have supported the viability of CO2 brines systems, it is

equally important to describe additional elements that provide increased efficiency

and plant safety.

Pump Package

While suction and discharge pump pressure can be read via pressure transmitters,

the use of differential pressure switches could play an important role in protecting

CO2 pumps. Mechanical pressure differential switches provide an augmented safety

consideration.

Regardingefficiency,VFDs(variablefrequencydrives)havealwaysplayedavitalrole

in increasing system efficiencies and they have become increasingly popular for two

reasons:

1. Energy savings.

2. Better liquid distribution in the evaporator coils. While energy savings cannot

typically justify the initial cost, liquid distribution cannot be overlooked. A

precondition for good distribution of refrigerant liquid is having a stable pressure

differential across the evaporators.

Pumpscontrolledbyfrequencyconverterscanensurethatpressureiskeptatstable

level under all load conditions.

Figure 8 depicts a comprehensive CO2 pump package. Beside the higher pressure

rating desired for the package the layout is fairly similar to regular ammonia pump

packages.




Pressure Control

Design pressure is one of the key factors of CO2 systems. CO2 brine systems,

especially those with defrost strategies other than hot gas, could be easily designed

for a maximum working pressure of 580 psig (40 bar). Therefore, most of the

standard off-the-shelve components could be used to build a safe system.

It is important to observe and anticipate standstill pressure conditions. One of the

alternatives is the installation of a small back-up unit to cool the re-circulator package

while the main power is down. This is a technique also used in CO2 cascade systems.

The capacity of such a unit is about 0.5 percent of the total plant capacity.

Another precautionary measure is to install check valves in parallel to all return

stop valves from the evaporators. In the event that the evaporators are shut down

for maintenance, all excessive pressure would be vented back to the main system

through the check valves.

Defrost Strategy

The most common ways of defrosting pumped CO2 systems are:

• Electricdefrost.

It is the simplest but often the least efficient method for defrosting air units.

It could be considered when the working temperature of the fluid is close to

freezing point. A thorough and complete analysis on defrost methods are available

illustrating that defrost efficiency ratio for electric defrost of a 100 kW evaporator

isabout53.8percentcomparedto226percentforhotgasdefrost.(Pearson,2006)

• HotGasDefrost.

It can be carried out either via a capable compressor that handles high discharge

pressures or a separate smaller compressor dedicated to generate hot gas. Figure




9 provides a schematic of evaporators with hot gas defrost. A high pressure

regulator is used to control the defrost pressure. After the defrost sequence is

finished,itisimportanttoverifytheMODP(maximumopeningdifferential

pressure) of the suction valve. Depending on the sizes and the design, a small

solenoid can be used to help equalize the pressure.

As CO2 brine systems are typically designed as oil-free, special care is required to

return the oil back to the defrost compressor.

• WaterDefrost.

In some cases (especially in rooms with temperatures above 32°F) evaporators

can be defrosted using sprayed water. There have been some available

technologies used to defrost low temperature evaporators.

• HotBrineDefrost.

This method requires that evaporators have a double coil (CO2 coil and brine coil)

which increases first cost. However, the defrost strategy is quite simple. The brine

could be heated by the NH3 condenser gas.

Valve Sizing Criteria

TherearenumerouspapersincludedintheIIARCarbonDioxideHandbookthat

addresses pipe sizing of liquid, wet return, compressor discharge, etc. The criteria

and methodology are consistent with best practices in terms of velocities and

pressure drop per 100ft of piping.

However, valves selection requires a slightly different approach to optimize flow and

achieve a desired temperature and/or pressure control.




The emergence of off-the-shelf evaporators, valve stations as well as motorized valves

with standard ratings of 52 bar (754 psig) simplifies the design and selection of

components to be used in both ammonia and CO2 sides.

As the pressure rating is no longer an obstacle in systems furnished with hot gas

defrost components, the next step is to select the appropriate port size to fulfill the

requirements.

Wet Return Lines

Control valves in wet return lines are subject to operating pressures which may vary

from 9 bar (132 psig) for low temperature units up to 28 bar (408 psig) for medium

temperature units. Therefore, 40 bar (600 psig) ratings were enough when using

electric or water defrost.

As a number of designers look at hot gas defrost, the wet return valves may be

subject to 44 bar or 655 psig if completing a defrost sequence at +10°C (50°F).

At first glance, a standard valve rated at 52 bar (754 psig) will be a great choice to

operate as a solenoid valve or control valve. However, the designer needs to look at

themaximumoperatingdifferentialpressureofthevalve(MOPD)aswell.Typically,

smallervalvesupto1.1/4"aredesignedtooperateatanMOPDof52baraswell.

Larger valves are normally rated to operate with lower differential pressures between

20 bar (290 psig) to 40 bar (600 psig).

This is important for a valve to be able to open after the defrost sequence is

completed and the evaporator returns to refrigeration mode. If the unit operates

at –40°F, then the wet suction valve would need to be capable of opening with a

differential pressure of 500 psig.




Hence, a pressure equalization process would be preferred by using a small by-pass

solenoid valve.

In terms of the actual port size of the valve, the use of 0.25°F to 0.5°F temperature

drop may be acceptable. The difference is quite noticeable when compared to an NH3

re-circulated evaporator.

Software widely used in the industry can be used to make the selection and or

permutation to obtain the desired port size (Dircalc v 1.20).

Fora50TRevaporatoroperatingat–40°F:

NH3 suction valve with a recirculation rate 3:1: would require a 5" motorized valve to

keep a pressure drop of 0.25 psig (temp drop: 0.8°F).

A CO2 suction motorized valve with a re-circulation rate of 1.5:1 would require a 2"

motorized valve with a pressure drop of 2 psig and a temp drop of 0.7°F.

As the CO2 system runs a pressure of 132 psig at those conditions, 2 psig does not

seem to be a critical number to overcome.

Last but not least, safety consideration during a power failure is necessary to take

into account.

There are, however, motorized valves that can be programmed to fully open or to

positionontoaspecificopeningdegreewhenconnectedtoUPSsystems.Thatwould

ensure the system would not trap liquid should a power failure occurs.




Pumped Liquid Lines.

Once again the pressure rating plays an important role in the selection process. 52

bar (754 psig) would be the requirement for the valve station as the check valve and

hand regulating valve (HEV) would be under 655 psig during the defrost sequence.

The liquid solenoid valve would probably determine the port size required and the

connection size would be determined by the velocity to match a recommended 3.28

ft/s (1 m/s).

Continuing with the example above, a 1.1/4" (DN32) connection size is good

choice. The port size of the solenoid valve is chosen to maintain the p-band of the

valve (pressure drop at which the valve is fully open). Though there are a number

of solenoid valves featuring zero pressure drop to open, it is important not to

exceed 2 psig or 3 psig as a general rule. A 1" (DN25) port solenoid valve would be

recommended.

Defrost Control Valve.

Portsizeselectionofthevalveiscrucialtoavoidundesiredchatteringindefrosting

CO2 evaporators. A defrost regulator with a Kv value of 3.5 m3/h(Cv:4.1USgal/min)

orslightlyhigherwouldbeenoughtodefrosta50TRunit.

Conclusions

There are different considerations to make when selecting the best choice for new

plants and additions to existing plants. It is clear that the use of CO2 in industrial

refrigeration has become a more viable option as current technologies make first cost

stand comparable to traditional systems.




However, looking beyond the first cost aspect and with an ever increasing focus on

plant safety, product quality and energy efficiency in industrial refrigeration systems,

CO2 brine systems make a serious case scoring high on each of those requirements

compared to water based solutions.

Dramatically reduced power consumption from the primary refrigerant and the

pumping power required can contribute to increasing the longevity of the equipment.

In turn, it brings another incentive to choose CO2 brine for medium and high

temperature systems in sensitive process areas due to the simplicity to operate the

plant.

Nevertheless, it is important to build the right eco-system to substantiate the benefits

of CO2 vs. glycol systems. Implementation of appropriate control strategies, loading

profile, maintaining zero moisture in the system, adopting the right components and

setting up safety management systems can help realize the great potential of CO2 to

end users throughout the industry.




References

Handschuh,R.2008.DesignCriteriaforCO2Evaporators.Guntner,Paperfor

Proklima–NaturalRefrigerants–SustainableOzoneandClimateFriendly

Alternatives to HCFCs.

Stenhede, C. 2007. Heat Exchangers in Carbon Dioxide Systems

Danfoss.PackCalculationIIandCO2 Brine Calculator 3.91 (computation program)

Pearson,A.2005.EvaporatorPerformanceinCarbonDioxideSystems.IIAR

conference and exhibition.

Vestergaard, N. 2004. CO2inSubcriticalRefrigerationSystems.IIARconferenceand

exhibition

Pearson,A.2006.DefrostOptionsforCarbonDioxideSystems.IIARconferenceand

exhibition.

Danfoss. Dircalc v1.20 (computational program)

Hinde, D. Zha, S. and Lan, Lin. 2009. Carbon Dioxide in North American

Supermarkets. Ashrae Journal

Mikhailov, A. Heiningen, K. and Kortstee, J. 2011. CO2 Secondary Coolant Systems:

Energy Efficiency and Control Strategy Considerations.

Berends,2006.NaturalWorkingFluidsinArtificialSkatingRinks.7thIIRGustav

Lorentzen Conference on Natural Working Fluids–Norway




Table 1: The results of pumping power required for various brines.

Temp. C 0 –5 –10 –15CO2/EthylenGlycol 9.0% 8.0% 7.1% 6.7%CO2/PropyleneGlycol 9.2% 8.1% 7.4% 6.1%CO2/CaCl2 7.5% 7.1% 6.3% 6.0%

Temp. C 0 -5 -10 -15CO2/Ethylen Glycol 9.0% 8.0% 7.1% 6.7%CO2/Propylene Glycol 9.2% 8.1% 7.4% 6.1%CO2/CaCl2 7.5% 7.1% 6.3% 6.0%

0.0% 1.0% 2.0% 3.0% 4.0% 5.0% 6.0% 7.0% 8.0% 9.0%

10.0%

0 -5 -10 -15

deg. C

CO2/Ethylen Glycol CO2/Propylene Glycol CO2/CaCl2




Table 2: The variables and assumptions made to calculate the total daily energy consumption.

CO2 Ethylene Glycol UnitsAir temperature, tair 5 5 °CCooling power, Qo 750 750 kWCirculation rate, n 1.5 1 –Temp. dif. in evap., dtevap 5 7 KTemp.dif.inPHE,dtPHE 4 5 KBrine temp. dif., tout - tin 0 4 KEvaporating temp., to –4 –7 °CAdditional heat gains, kq 5% 7% %Additional heat gains, Qadd 37.5 50.6 kWPumpheadpressure,Hpump 2.5 2.5 BarPumppowercons.,Ppump 1.6 16.6 kWAdjusted cooling power, Qo,ad 789.1 815.6 kWCompr.powercons.,Pcomp 170.8 196.2 kWWorking hours, daily 18.0 hTotal install energy cons. 172.4 212.8 kWTotal daily energy cons. 3,112.9 3,930.3 kW*hEnergy savings 21% %




Table 3: Side to side comparison for both plants.

Location North Italy The NetherlandsStorage type Fruit FruitAir temperature °C –2 to +6 0,5Suction temperature, °C –13 –10,5Brine Glycol CO2Media temperaturem °C (for glycol – in/out) 0.5 –7,5 (not optimized)temperature difference in cascade heat

exchanger difference, K

5 to 7 3 to 4

Cooler temperature difference, K 8 to 12 8Lighting, estimated from the total load 10% 5%Total fans installed, kW 74,1 453,3 (fans are

running only 20%

of the time)Total consumed, kW h (measured figure) 1.300.000 2.700.000 Total compressors and pumps, kW h

(calculated, excluding lights and excluding fans)

576.607 1.245.025

Regioncorrectedconsumption,kWh,by

PackCalculationIIsoftware,accordingtothe

assumptions above, everything adjusted to the

Netherlands climate

518.000 1.245.025

Average consumption, kW (calculated by

PackCalculationIIsoftwaretofitthecorrected

consumption)

188 660

PerunitofcoolingconsumptionkWh/kW

cooling (corrected consumption / average

consumption)

2.755 1.886

Difference 32%




Figu

re 1

: A p

umpe

d am

mon

ia v

alve

arr

ange

men

t fo

r m

ediu

m a

nd/o

r

hi

gh t

empe

ratu

re u

nits

.




Figu

re 2

: A h

ybri

d re

frig

eran

t/gl

ycol

sys

tem

.




Figure 3: A conventional CO2 system showing the medium temperature loop.




Figu

re 4

: A s

impl

ified

CO

2 sy

stem

and

the

cor

rela

tion

of t

he d

iffer

ent

cond

itio

ns o

f the

refr

iger

ant

in t

he lo

g P

-H d

iagr

am.




Figure 5: Dramatic reduction in pipe sizes for both supply and return of medium and low temperature systems.




Figu

re 6

: Whe

n th

e de

sire

d te

mpe

ratu

re is

rea

ched

, a 7

54

psig

pre

ssur

e ra

ted

valv

e

st

atio

n ca

pabl

e of

pul

se w

idth

mod

ulat

ion

redu

ces

the

“ope

ning

” de

gree

by

redu

cing

the

tim

e of

liqu

id in

ject

ion

base

d on

a fi

xed

cycl

e.




Figu

re 7

: A g

raph

of t

he te

mpe

ratu

re v

aria

tions

whe

n us

ing

puls

e m

odul

atio

n te

chno

logy

.




Figure 8: A comprehensive CO2 pump package.




Figure 9: A schematic of evaporators with hot gas defrost.




Notes:


Notes:




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