project 2: co in supermarket refrigeration - kth · the project “co2 in supermarket...

Effektivare butikskyla

Project 2: CO2 in Supermarket Refrigeration

Samer Sawalha

Arash Suleymani

Jörgen Rogstam

2006

CO2 Project Report-Phase I

1. SUMMARY The project “CO2 in Supermarket Refrigeration” is collaboration among IUC, KTH/Applied Thermodynamic and Refrigeration division, Ahlsell, Huurre, AGA, WICA and ICA. The project is financed by Energimyndigheten. The objective of this project is to develop, test, and evaluate an energy efficient supermarket system working with CO2 as the refrigerant. Based on the experience in designing the system, running and evaluating it, modifications should be applied in order to conclude an efficient optimized CO2 system for a medium size supermarket in Sweden. Emphasize is on using environmentally friendly refrigerants and the choice was to use natural fluids. A refrigeration system solution for a medium size Swedish supermarket has been built in IUC laboratory in Katrineholm. The system is equipped with extensive instrumentations to collect data and perform online diagnosis. Several variations of the system solution are applied for validation and possible modifications. In this report we present the system under investigation and some of the experimental results that have been obtained under the project period. Overall system validation and evaluations of the main components are described.

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2. FORWARD This report is a product of the research program Effektivare Kyla which is organized and managed by IUC in cooperation with KTH. The program is a shared financing between Energimyndigheten (STEM) and the involved companies. We would like to extend our appreciation for the positive spirit of cooperation from all the project partners. We thank everyone who worked for realizing the idea of the project and maintain its activities. Samer Sawalha Arash Soleymani Jörgen Rogstam January 2006

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3. TABLE OF CONTENT

1. Summary ........................................................................................................................ 2 2. Forward .......................................................................................................................... 3 3. Table of content.............................................................................................................. 4 4. Introduction .................................................................................................................... 5 5. Objectives....................................................................................................................... 7 6. The case study ................................................................................................................ 8

6.1 System Requirements and Boundaries....................................................................................... 8 6.2 The System Solution .................................................................................................................. 8

7. The experimental rig .................................................................................................... 10 8. Ovearall system analysis .............................................................................................. 12

8.1 Load Measurements ................................................................................................................. 12 8.2 Energy Balance Test ................................................................................................................ 12

Cascade condenser cooling capacity ........................................................................................ 15 Ammonia compressor capacity and efficiencies ...................................................................... 19 CO2 compressor capacity and efficiencies ............................................................................... 24

8.3 System Efficiency .................................................................................................................... 29 Low stage COP......................................................................................................................... 29 High stage COP........................................................................................................................ 30 Total COP................................................................................................................................. 30

9. System variations and experimental results ................................................................. 32 9.1 Pump Circulation Ratio and Gravity Circulation..................................................................... 32 9.2 Cascade Condenser .................................................................................................................. 37 9.3 Flashing Gas in Liquid Lines ................................................................................................... 40 9.4 Freezing Cabinets Control........................................................................................................ 41 9.5 Safety Tests .............................................................................................................................. 50 9.6 Flooded Versus Direct Expansion Evaporation ....................................................................... 50

10. Theoretical model..................................................................................................... 51 10.1 Model Description.................................................................................................................. 51 10.2 Overview of Calculations and Assumptions .......................................................................... 51 10.3 Design Calculations ............................................................................................................... 52 10.4 Some Simulation Results ....................................................................................................... 53

11. Conclusions and future work.................................................................................... 54 12. References ................................................................................................................ 56

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4. INTRODUCTION In the early stages after the revival of CO2 as a refrigerant it was applied in low temperature applications in supermarkets as a secondary refrigerant in indirect systems. The usage of CO2 in cascade and trans-critical systems for this application has been suggested along with the indirect system, but the limitations for the application of the cascade and trans-critical systems have been related to the scarcity of components that can efficiently handle CO2. Also there have been many unanswered questions related to how to handle the highly pressurized system and how safe it is to deal with CO2 in this application. In the indirect system it was possible to use conventional components to handle CO2; this is mainly due to the low operating pressure in the indirect loop at low temperature applications, 12 bars at -35°C. The pumping power needed for CO2 in the indirect system is very small compared to conventional brine systems due to the small volume flow rate and pressure drop of CO2 in the circuit. The small volume flow rate is a result of the phase changing process on the CO2 side, which also contributes to having small pressure drop in the pipes and heat exchangers. Gaining experience and confidence in working with CO2 in indirect systems combined with the knowledge that is gained through extensive research work on CO2 in mobile air conditioning and hot water heat pumps brought the attention of the industry for the necessity of producing components which are specially designed, or modified, to handle CO2. As a result, cascade and trans-critical systems became reasonably applicable. In supermarket applications the difference between evaporating and condensing temperatures is large, therefore, the cascade or other two-stage systems become favorable and they are well adaptable for the two-temperature level requirement for chilled and frozen products in the supermarket. The indirect system requires an additional heat exchanger (primary refrigerant evaporator/CO2 condenser), which implies that there is an additional temperature difference across the heat exchanger and the resulting evaporating temperature will be lower than if a direct expansion had been performed. In cascade or multi-stage CO2 systems, CO2 evaporates directly in the evaporators of the display cases, which minimizes the required temperature lift and reduces the energy consumption. In the direct expansion solution, the CO2 pump is not required; despite the fact that the power consumption of the pump in the CO2 indirect solution is generally very small compared to conventional brine pumps and relative to the total power consumption of the entire system, still the elimination of the pump is advantageous to reduce the installation cost. Practically the pumping power usually is higher than necessary as it is difficult to find CO2 pumps to match medium capacity systems and as the pumps usually are larger than needed a bypass line is introduced to obtain the required CO2 flow rate in the evaporators. The low critical point for CO2 of 31ºC implies that it will operate with better theoretical COP between low temperature ranges further below the critical point compared to high stage conditions. Therefore, the application of CO2 in the low stage of a cascade system yields a reasonable theoretical COP compared to other refrigerants, see Figure 1a and Figure 1b. In Figure 1b, the COP values above the critical point are obtained at optimum pressure on the high pressure side.

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(a) (b) Figure 1 COP of CO2 compared to other main refrigerants in low stage (a) and in high stage (b) operations

Several factors contribute to improve the COP of CO2. The favorable thermophysical properties of CO2 results in low pressure and temperature drops in the system. From the heat transfer point of view, the low surface tension will make boiling easier and therefore will improve the heat transfer. Also, due to the low pressure drop, (Zhao, Molki et al. 2000), the components of the system will be smaller while the mass flow rate of the refrigerant will be comparable to R404A, R22, R502 and R134a refrigerants, which will result in high mass flux of CO2 in the heat exchangers. Another improvement to the COP comes from the improved volumetric efficiency of the CO2 compressor compared to conventional refrigerants; this is due to the lower pressure ratio across the CO2 compressor.

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5. OBJECTIVES The general experience from the installed CO2 refrigeration systems in supermarkets is positive and usually systems’ performances are comparable to similar traditional solutions. The observations made on the CO2 systems are usually obtained from real installations in supermarkets or industrial plants and they are usually compared to installed conventional systems or related to experiences in design, installation and running of previous installations of conventional systems. In order to advance in developing CO2 system solutions for supermarket refrigeration there is a need for detailed evaluation of the performance of the different solutions. The main objective of this project is to develop, test, and evaluate an energy efficient supermarket system working with CO2 as the refrigerant, emphasize is on using environmentally friendly refrigerants and the choice was to use natural fluids. A laboratory environment allows control over the boundary conditions of system and provides flexibility for modifications. Investigations will focus in first place on overall system evaluation. Detailed analysis of the main components of the system is also an important possibility provided in laboratory environment. Comparison of different cycle modifications will allow optimization of the system for the most efficient solution arrangement, operating conditions and control strategies. The experimental model when evaluated against a theoretical one will prove how close the system and refrigerant behave close to predictions. It is essential to build a theoretical model to facilitate the design and sizing of the system’s components. The agreement between the two models should give confidence to apply modifications to the theoretical model and test the efficiency of the system and point out the components with highest energetic inefficiencies. Based on the experience in designing the system, running and evaluating its modifications, it will be possible to conclude the most efficient and optimized CO2 system for a medium size supermarket in Sweden. It will be possible to point out the components in the system that have low performance and that will direct the research in the future to improve the efficiency of such systems.

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6. THE CASE STUDY 6.1 System Requirements and Boundaries The system solution under investigation has been chosen to be applicable to a medium size supermarket installation in Sweden. The refrigeration system in the supermarkets in Sweden usually operates to satisfy the evaporating temperatures to maintain products at two temperature levels, around +2ºC for cold food and -18ºC for frozen products. Despite the low ambient temperature in Sweden the condensing temperature is usually kept constant around the year at a value of about 40 ºC. The cooling capacities of medium size supermarkets are typically around 50 kW for freezing and 150 kW for cooling at the medium temperature level. This estimate is based on contacts with major installers of supermarket systems in Sweden. Accordingly, the system has been designed to operate between the temperature boundaries mentioned above and to provide a cooling capacity which is scaled down while trying to keep a load ratio of about 3. The low temperature side has a rated capacity of 7.4 kW and the medium temperature side was designed to have a capacity of 20 kW. 6.2 The System Solution In the choice of the CO2 system the aim was to develop an efficient system with good cooling performance, safe and in accordance with the regulations on the use and release of refrigerants. From environmental point of view, the amount of synthetic refrigerants that can be used to fill the system is limited and the high taxation to prevent leakage makes it expensive to use. For natural refrigerants such as ammonia and propane there is always the safety concern when used in applications where people might be exposed to a leakage accident. CO2 is considered as a relatively safe refrigerant and classified in group A1 according to ASHRAE Handbook-Fundamentals (1997). CO2 gas that is used in refrigeration is a by-product of the chemical industry and its use in the refrigeration system can be considered as a delayed step before its unavoidable release to the environment. Therefore, CO2 becomes an interesting solution as a refrigerant from environmental and safety points of view especially in supermarket refrigeration systems where large quantities of refrigerant are required and direct contact, in case of leakage accident, with large number of people might occur. The system that has been chosen is a cascade system with NH3 at the high stage and CO2 at the low stage, at the medium temperature level CO2 is pumped to provide the required cooling load. Figure 2 is a schematic diagram of the CO2 circuits in the system. The usage of the cascade system offers the possibility of utilizing two different refrigerants where each refrigerant is selected to fit the operating range. Using NH3 in the high stage means that it will be easy to deal with a leakage accident as ammonia can only leak into the machine room which should be equipped with proper safety devices. Using CO2 in the low stage results in reasonable operating pressure levels in the CO2 circuit, 28 bars at -8°C. The favorable pressure drop characteristics of CO2 suits this application where long distribution lines are usually needed. Also this implies that the size of the distribution lines is also smaller than for other refrigerants which reduces the cost of the piping system.

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The evaporators at the medium temperature stage are flooded with CO2 which is circulated via a pump; this is expected to produce better performance due to the good heat transfer characteristics of the completely wetted evaporator and therefore the evaporator temperature will be higher than if a direct expansion concept has been used.

Figure 2 Schematic diagram of NH3/CO2 cascade system with CO2 at the medium temperature level

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7. THE EXPERIMENTAL RIG The 20 kW design cooling capacity at the medium temperature level is divided over two display cabinets with 5 kW each and the other 10 kW’s are supplied by an electric heater in what is referred to as a load simulator. The installed electric heater at the medium temperature level managed to provide a maximum of 6.6 kW, which makes the maximum total load at the medium temperature level reduces to 16.6 kW. The electric heater load can provide three load steps of 2.2, 4.4 and 6.6 kW. On the deep freeze side, the load is divided over two freezers with 2.5 kW each and the electric heater can provide a maximum load of 3 kW. The electric heater load can be provided on three equal load steps in a similar way to the medium temperature load simulator. The maximum cooling capacity of the compressor is 7.4 kW. The freezers are equipped with electronic expansion valves. The compressor is a Copeland scroll type with operating temperatures between -37°C and -8°C and a displacement of 4.1 m3/h. The accumulation tank has a capacity to contain 180 L of CO2 and is equipped with an electronic level indicator. It can stand a pressure up to 40 bars which corresponds to an operating temperature of about 6 °C. The system is equipped with a safety release valve that is triggered when the pressure in the system reaches 38 bars. To avoid the opening of the release valve and the loss of significant charge from the system, a bleed valve is installed which opens for periodical release of CO2 at lower pressure than the set value for the release valve, 35 bars, so the pressure in the system will be reduced. If the pressure increase in the system is higher than the rate that the bleed valve can handle, then the release valve will open and release the system’s charge. The CO2 pump that is used is a hermitic one with capacity higher than the highest circulation rate desired; therefore a by-pass is used to reduce the flow rate pumped into the medium temperature circuit. About 1.5 meters head over the pump is respected to prevent cavitation. The ammonia unit uses a Bock reciprocating compressor with displacement of 40.5 m3/h; it can run at 50% reduced capacity by unloading half of its cylinders. Heat is removed from the ammonia evaporator via a thermosyphon loop which required a certain height of the unit. The capacity control of both compressors is achieved by a frequency converter. The cascade condenser is a plate type heat exchanger that is specially selected to handle the pressure difference that will exist between CO2 and ammonia, at -8 °C CO2 will have about 28 bars while ammonia will have a pressure of about 2.7 bars at -12 °C. The medium temperature display cabinets are defrosted using the conventional electric defrost method. The deep freeze cabinets will be defrosted using hot gas defrost by passing the hot gas through the cabinet, since the condensing temperature of the CO2 at the compressor discharge pressure will be low then heating of the evaporators will be achieved via the sensible heat of the hot gas. Figure 3 is a detailed schematic of the test rig with most of the measuring points indicated. As can be seen in the schematic several by-pass lines and the high number of valves indicate the

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possible variations in the system which will be used for testing and modifications. The load simulators with the electric heaters can be seen in the diagram in the medium and low temperature circuits. The electric heater provides heat to a brine loop which exchanges the heat with the refrigerant in a plate heat exchanger. The schematic shows that one of the freezers is electrically defrosted while on the rig both freezers are equipped with the hot gas defrost.

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Figure 3 Detailed schematic diagram of the NH3/CO2 cascade system test rig

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8. OVEARALL SYSTEM ANALYSIS The system under investigation is a scaled down real installation where the main discussion is weather or not this system is a suitable replacement for traditional technologies. In order to provide answers about the current system solution, it is important to perform the overall analysis of the system where the capacities are properly measured, the energy balance is verified, and the system’s efficiencies are calculated according to the measurements. This is an important step since some of the measurements are based on the components data and planned experiments depend on the accuracy of measuring cooling loads and capacities of some of the main components. 8.1 Load Measurements The two compressors are used to determine the mass flow rate of the refrigerant which will then be used to calculate the cooling capacities in the corresponding circuits. The compressor manufacturer data have been used as guidelines for the calculations which are based on knowing the geometry and the efficiencies of the compressors at certain operation conditions. Measuring the rotational speed of the compressor, and the temperatures and pressures around it gives all the data needed to calculate the mass flow of the refrigerant. Consequently, it will be possible to calculate the energy consumption of the compressor and the cooling capacities of the evaporators/cabinets. At the return line of the medium temperature level the flow is a two phase one, therefore it is not possible to calculate the load at the medium temperature by measuring the mass flow of the refrigerant. By calculating the cooling capacity at the cascade condenser and for the low stage cabinets it will be possible to calculate the total load at the medium temperature level. 8.2 Energy Balance Test The simulators at the medium and low temperature levels provide a fixed known cooling capacity via the electric heaters which can be used to verify the method of calculating the cooling capacity at the medium and low temperature levels using the compressors manufacturers’ data. The medium temperature simulator provides a maximum of 6.6 kW, and the low temperature simulator provides a maximum of 3 kW. The two simulators can be switched to 1/3 and 2/3 of its capacity. The system is run with only the simulators on in addition to the CO2 pump and compressor. The blue line in Figure 4 shows the active lines and components in this test.

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Figure 4 Schematic of the active lines and components in the energy balance test

The electric power consumption of the CO2 pump is measured and it varied around the average value of 0.85 kW. The power consumption of the CO2 compressor was measured by an electric meter instead of using the manufacturers’ data, this is due to the fact that the isentropic and volumetric efficiencies were much less than the provided data. Calculating the mass flow using the compressor data resulted in much higher cooling capacities than the 3 kW provided by the simulator. The compressor was running at a constant rate and the power consumption was measured to be around 1.7 kW. The system is operated around 32ºC for condensing ammonia, -26ºC for freezers, and a medium temperature of about -9ºC, a plot for the temperatures of the boundary conditions of the system during the test period is presented in Figure 5.

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Evaporating and Condensing Temperature

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Figure 5 Plot of the temperatures at the system boundaries during the energy balance test period

The system is first run with only the medium temperature load (simulator and pump); region 1 shown in Figure 6, and then the low temperature circuit is switched on where the low temperature simulator and the compressor power capacities are added to the load at the medium temperature, region 2. The ammonia compressor run at reduced volume of 50% by unloading two cylinders and this is the volume that have been used along the test except in the regions 2.1 and 2.3 where the compressor was switched to full volume. Running the compressor at full volume resulted in occasional stop start operation where the load seemed to be lower than the lowest load to maintain continuous operation of the compressor while switching to half of the compressor effective volume region 2.2 showed that the compressor was running at full speed and was not able to reduce the pressure in the tank to the set value. In order to maintain continuous operation of the compressor the load at the medium stage was reduced to 2/3 of the total capacity with half of the ammonia compressor volume, region 3. Further reduction of 1/3 on the medium temperature was performed in region 4 after which the simulator was switched off. The CO2 pump at the medium temperature level was kept running in region 5 and then switched off in region 6. In region 7 the low stage was switched off and the system was running with only the CO2 pump on.

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Energy Balance

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Figure 6 Energy balance with fixed value of volumetric efficiency

Running the system at different capacities aims at verifying that the cooling capacity calculated by the ammonia compressor matches different capacities in the system. Cascade condenser cooling capacity The ammonia mass flow that is passing through the cascade condenser is calculated using the ammonia compressor data, the following equation is used to calculate the refrigerant mass flow of the compressor:

inss Vm ρη ⋅⋅= && (1) Where is the swept volume flow in msV& 3/s, inρ is the density of the refrigerant (kg/m3) at the inlet of the compressor, sη is the volumetric efficiency of the compressor. The compressor has a displacement ( ) of 40.5 msrV& 3/hr at rated speed ( ) of 1450. Swept volume flow in mrn 3/s at a given speed can then be calculated using the relation:

36001

⋅⋅=r

srs nnVV && (2)

Where is the compressor speed in RPM (1/min). n The volumetric efficiency of the compressor was extracted from BOCK software by running the software at different operating conditions, for the same operating conditions the mass flow is calculated for an ideal compressor, the ratio of the two values is the volumetric efficiency. An average value for different operating conditions was found to be about 85%.

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Measuring the compressor speed, the temperature and pressure at the inlet of the compressor provide the required data to calculate the mass flow. The conditions before and after the cascade condenser are also measured and therefore the cooling capacity is calculated using the relation:

dhmQ ⋅= && (3) Where dh is the enthalpy difference across the heat exchanger. In Figure 6, the cooling capacity of the cascade condenser is plotted and the difference between this capacity and the provided load is also plotted as dQ which is also presented with average values (dQaverage) over the different operating regions. Excluding region 2 where it was hard to reach the set point and load peaks have been observed at transition time, the difference in load which varies between 1 and 1.8 could be explained as heat losses in the system and deviation in the volumetric efficiency from the estimated value from BOCK’s software. The volumetric efficiency value that was used in the calculations presented in Figure 6 is assumed to be constant, which is not the case in practice and it will vary with the pressure ratio. In order to correlate the volumetric efficiency to the operating conditions, a relation suggested by Pierre (1982) for “good” ammonia reciprocating compressor is used to calculate the volumetric efficiency as follows:

⎟⎟⎠

⎞⎜⎜⎝

⎛⋅−⋅=

2

1063.0exp02.1PP

sη (4)

Where 2

1

PP is the pressure ratio.

Using the volumetric efficiency calculated from equation 4 to calculate the mass flow and the cascade condenser capacities yields the results in Figure 7. It can be seen from the figure that the difference in cooling capacity is reduced due to the reduced value of the volumetric efficiency and the value varied around 1 kW with smaller deviation around the average value compared to the trend in Figure 6.

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Energy Balance: calculated volumetric effeciency

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Figure 7 Energy balance with calculated value of volumetric efficiency

If the heat sink into the system is neglected assuming that the difference in the load is due to lower volumetric efficiency than the calculated one then it will be possible to calculate how much the “actual” volumetric efficiency of the compressor should be by using the known load as the input value to calculate the mass flow of refrigerant from equation 3. Consequently, the volumetric efficiency can be calculated using equation 1. The plot in Figure 8 shows the volumetric efficiency calculated using the relation in equation 4 and the “actual” volumetric efficiency. As can be seen in the plot the difference between the calculated and actual values is about 10% less for the actual efficiency. Therefore it will be possible to reduce the value calculated in equation 4 by 10% in order to adjust the calculated volumetric efficiency to be closer to the actual value. The third plot in Figure 8 is the adjusted value of the volumetric efficiency.

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Ammonia compressor volumetric effeciency

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Figure 8 Volumetric efficiency of the ammonia compressor

Using the adjusted value for the volumetric efficiency brings the average value for the difference in cooling capacities closer to zero as can be seen in the figure below.

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Energy Balance: adjusted volumetric effeciency

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Figure 9 Energy balance with adjustment to the calculated value of volumetric efficiency

Ammonia compressor capacity and efficiencies The ammonia compressor is used to estimate the cooling capacity of the cascade compressor and the therefore it is important to evaluate its performance and capacities. Pierre (1982) suggests a relation to estimate the isentropic efficiency ( kη ) for the same compressor relating it to the volumetric efficiency in equation XX.

⎟⎟⎠

⎞⎜⎜⎝

⎛+⋅−= 97.169.1exp

2

1

TT

k

s

ηη (5)

The temperature ratio is the absolute temperatures, in Kelvin, of condensation and evaporation corresponding to exist and inlet compressor pressure. Figure 10 shows the calculated and measured values of the isentropic efficiency. Using BOCK software the average isentropic efficiency obtained is about 76% over a range of different operating conditions, which is close to the calculated one, around 78%. The higher value of the measured efficiency may have to do with position of the temperature sensor at the discharge line; it may sense lower temperature value than the actual one.

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Ammonia Compressor Isentropic Effeciency

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Figure 10 Measured and calculated isentropic efficiencies of the ammonia compressor

Another method that can be used to calculate the cooling capacity is to measure the electric power consumption of the compressor and the enthalpy difference across the compressor is determined by measuring the pressures and temperatures across it. Certain electric motor efficiency and heat losses to the environment should be assumed in order to calculate the shaft power according to the following equation:

elelthermalShaft EE && ⋅⋅= ηη (6) Referring to Climate Check a 7% of thermal losses ( %93=thermalη ) is usually assumed. Usually the shaft power is considered to be the power that is provided to the refrigerant and the mass flow can be calculated according to the equation below, the mass flow is then used to calculate the cooling capacity in equation 3.

compshaft dhmE ⋅= && (7) Figure 11 shows typical efficiency values for well performing electric motor related to the shaft power Granryd et el. (2003). The shaft power in the experiment is ranging between 1.8 and 3.3 kW, as seen in Figure 12, according to the figure below the electric motor efficiency is estimated to be around 80% at the rated speed.

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Figure 11 Typical values of electric motor efficiency versus the motor shaft power Granryd et el. (2003)

In the case when the compressor is running at reduced cylinder capacity then there will be additional losses attached to running two non-productive cylinders, in this analysis these losses are included in the electric motor efficiency, therefore it will be lower than the estimated value. The input electrical power is measured and plotted in Figure 12 along with the shaft power calculated from the refrigerant side; the adjusted shaft power is the power that is calculated based on adjusted volumetric efficiency presented in Figure 8.

Ammonia Compressor Power Consumption

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Figure 12 Calculated ammonia compressor shaft power and measured electric motor efficiency

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Using equation 6 to calculate the efficiency of the electric motor produces the results in the plot of Figure 13. The electric motor efficiency value is dependant on the motor speed which is also shown in the figure.

Electric Effeciency vs RPM

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12.1 2.2 2.3 3 4

Figure 13 Calculated electric motor efficiency for the energy balance test

In regions 1, 2.2, 3 and 4 the compressor is running at reduced capacity with two unloaded cylinders which will add some mechanical losses; these losses are included in the electric motor efficiency presented above. In addition to that in most of the regions the compressor was running at partial speed which reduces the efficiency of the electric motor. In the end part of region 2.3 the compressor was running at full cylinder capacity. Where the operation was close to being stable the efficiency increased due to less mechanical/friction losses, but still the compressor was running at partial speed which reduces the electric motor efficiency. In another test the compressor was running at higher cooling capacity with all the cylinders active. The motor speed was varying due to test conditions and the electric motor efficiency is calculated in a similar manner to the approach above. The plot in the figure below shows that the electric motor efficiency is higher than the values presented in Figure 13. This is mainly due to the fact that all the moving cylinders are producing work. It can be seen more clearly in the figure below how the efficiency changes with the speed of the motor. The trend of change is plotted in Figure 15 where it indicates that the efficiency tend to increase with motor speed. The results in the figure should not be read as quantitative; the aim of the plot is show the tendency of behavior.

22


Electric Effeciency vs RPM

40

50

60

70

80

90

100

14:16:48 14:52:48 15:28:48 16:04:48 16:40:48 17:16:48 17:52:48

%

800

1 300

1 800

2 300

2 800

RPM

Electirc Effeciency Electirc Effeciency Adjsuted Ammonia RPM

Figure 14 Calculated electric motor efficiency for ammonia compressor with all cylinders running

Electric Motor Efficiency vs RPM

40

50

60

70

80

90

100

110

120

800 900 1 000 1 100 1 200 1 300 1 400 1 500RPM (1/min)

Elec

tric

Effe

cien

cy (%

)

Figure 15 Calculated electric motor efficiency for ammonia compressor at different motor speeds

23


CO2 compressor capacity and efficiencies The same method that is used with ammonia unit to calculate the mass flow of refrigerant from the compressor manufacturer data was used for the CO2 compressor; the results showed a large deviation from the expected values, the resulting cooling capacity was much higher than the provided 3 kW load. This indicates that the actual volumetric efficiency of the compressor is lower than the 90% value used from the manufacturer data. Therefore, the CO2 compressor power that is used in the energy balance was measured by the electric energy meter; the average recorded value was 1.62 kW which can be considered as constant since the load at the low stage was constant and equal to the electric heater capacity of 3 kW. Knowing the capacity and the conditions around the freezer simulator the mass flow of refrigerant can be calculated and then can be used to calculate the power consumption of the CO2 compressor. This way of calculating the compressor power resulted in lower value of about 15% than the measured electric power consumption. This can be seen in the figure below.

CO2 Compressor Power Consumption

-0,20

0,00

0,20

0,40

0,60

0,80

1,00

1,20

1,40

1,60

1,80

13:55:12 14:24:00 14:52:48 15:21:36 15:50:24 16:19:12 16:48:00 17:16:48

kW

CalculatedIsentropic Eff=60%Average Measured

Figure 16 Calculated CO2 compressor shaft power and measured electric motor efficiency

The reason for the deviation may be partly due to the difference in the actual discharge temperature and the measured value. The point where the temperature is measured is very close to the compressor exit but due to the very high discharge temperature and the big difference with the room temperature it might be possible that some of the deviation is due to

24


this difference. In order to estimate how much the actual discharge temperature should be, the measured compressor power is used to calculate the enthalpy at the exit of the compressor and consequently the temperature can be calculated by using the discharge pressure. Figure 17 shows the measured and the expected actual value.

CO2 compressor discharge temperature

50

60

70

80

90

100

110

120

130

140

150

160

170

180

190

200

14:24:00 14:52:48 15:21:36 15:50:24 16:19:12 16:48:00 17:16:48

Tem

pera

ture

C

T_hg_calcT_hg_meas

Figure 17 Calculated and measured CO2 compressor discharge temperature

In order to evaluate the isentropic efficiency of the compressor the manufacturer’s data is compared to the measured ones. Figure 18 shows the curve from the manufacturer data where the isentropic efficiency at the running pressure ratio of about 1.7 results in an estimated isentropic efficiency of 60%, assuming that the curve does not drop sharply before the optimum point. Using this value to estimate the power consumption of the compressor in case of a “good” compressor results in the power consumption presented in Figure 16. It was not possible to reach the optimum condition for the pressure ratio due to the fact that reducing the evaporation pressure resulted in the discharge temperature increasing to very high levels.

25


Isentropic Efficiency

0,40

0,50

0,60

0,70

0,80

0,90

1,00

1,00 2,00 3,00 4,00 5,00 6,00

Pressure Ratio

Figure 18 Isentropic efficiency of the CO2 compressor at different pressure ratios

Using the calculated mass flow from the simulator side and knowing the compressor geometry and RPM it will be possible to calculate the volumetric efficiency that is presented in Figure 19 along with the measured isentropic efficiency which is calculated by measuring the pressures and temperatures around it.

26


CO2 compressor Isentropic and volumetric effeceincies

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35,00

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45,00

50,00

13:55:12 14:24:00 14:52:48 15:21:36 15:50:24 16:19:12 16:48:00 17:16:48

%

Isentropic effeceincyVolumetric effeceincy

Figure 19 Measured Isentropic and volumetric efficiencies of the CO2 compressor

It is evident that the volumetric and isentropic efficiencies of the compressor are low and this is due to the fact that the compressor have been operating with unfavorable conditions during early runs of the system. It has been notices that the valve pointed out in Figure 20 was leaking even when it was firmly closed. When the medium temperature circuit was running with the low stage off liquid CO2 leaked into the oil separator which resulted in bad lubrication for the compressor. This may have created damage in the compressor which was running with normal sound after the valve was replaced and provided the required cooling capacity but at a higher cost of energy consumption due to the low volumetric and isentropic efficiencies.

27


Minimum flow nozel

Safety pipe

p

p

M

M

t

t

pt

t

t

t

t

t

t

Oil drain

p


p

5 kW

t

tt p

p

El D

efrostE

l Defrost

6,6 kW el

ttdP

5 kW tdP

t

Leakage test point

Cha

rgin

g po

int

t

2.5 kW

t

El Defrost

dP

t

2.5 kW

Hot gas Defrost

t

3 kW

el t t

dP

p

t

dP

t

dP

t

Figure 20 Schematic diagram of the system shows the leaking line

28


8.3 System Efficiency The overall system efficiency is evaluated using the COP which relates the useful load to the work done to provide it. In this test for the energy balance verification the cooling capacities are known by running the system on electric heaters in the load simulators as the only load in the system. The power input to the system has been obtained by measuring the electric power consumption of the compressors and pump. This means that all the needed parameters to calculate the COP are available. Low stage COP The known supplied load to the low stage via the electric heater and the electric power consumption of the CO2 compressor is measured and the resulting COP is plotted in the figure below and denoted as actual value. The calculated COP in the figure is based on the power consumption that is calculated using the mass flow obtained from the simulator side. Assuming a good compressor that fulfils the manufacturer’s specifications with the estimated 60% isentropic efficiency then the power consumption of the compressor will be much lower with a high COP, as can be seen in the figure.

Low Stage COP

0,00

1,00

2,00

3,00

4,00

5,00

6,00

7,00

8,00

14:09:36 14:38:24 15:07:12 15:36:00 16:04:48 16:33:36 17:02:24

CO

P

Calculated, Eta_v=60%CalculatedActual

Figure 21 Calculated and measured low stage COP

With CO2 compressor that has volumetric and isentropic efficiencies closer to the design values the energy consumption of the compressor will be reduced and this will improve the low and total COP’s.

29


High stage COP In the case of the ammonia unit performance evaluation, the capacity of the cascade condenser is considered as the “useful” load which includes the medium, low stage, CO2 compressor, and pump capacities. The COP of the ammonia presented in the figure below is calculated by referring to the directly measured electric power consumption which is referred as “actual”. The “calculated” COP is the one that it is obtained by using the mass flow of refrigerant to calculate the compressor power; the mass flow is based on the calculated volumetric efficiency.

Ammonia Stage COP

0

1

2

3

4

5

6

7

8

9

10

12:28:48 13:04:48 13:40:48 14:16:48 14:52:48 15:28:48 16:04:48

CO

P

Useful cooling loadMeasuredCalculated

Figure 22 Calculated and measured ammonia unit COP

The measured COP’s of the ammonia unit are lower than expected and this is mainly due to running the system at partial load which is companioned with some mechanical and electric motor losses. At loads close to full capacity the ammonia compressor will have lower electric power consumption, consequently, high stage and total system COP’s are expected to improve. Total COP The total COP is calculated using two different values for the ammonia compressor power consumption which are discussed in the ammonia unit evaluation above.

30


Total COP

0

1

2

3

4

5

12:28:48 13:04:48 13:40:48 14:16:48 14:52:48 15:28:48 16:04:48

CO

P

4

5

6

7

8

9

10

CalculatedActualUseful cooling load

Figure 23 Calculated and measured total COP

Improving the efficiencies of the ammonia and CO2 compressors with favorable operating conditions on both stages and new CO2 compressor will reduce the power consumption and reduce the losses at partial load. This will improve the total COP of the system.

31


9. SYSTEM VARIATIONS AND EXPERIMENTAL RESULTS In practice, there are many possible variations within the applied CO2 systems, these variations should be investigated in order to modify the system’s design and optimize its cost and improve its performance. In this test rig many variations of the system have been made possible and the design took into consideration the need to modify and adjust some operating parameters. 9.1 Pump Circulation Ratio and Gravity Circulation At the medium temperature side the circulation ratio of the refrigerant can be varied so the pressure drop and heat transfer in the display cabinets and distribution lines can be investigated, low circulation ratios will result in low pressure drop but may influence the heat transfer in the cabinet. A by-pass line will be used to change the circulation ratio, as using a variable speed pump was impossible due to the fact that it was not possible to find a small pump that can cover the desired flow range. The pump can also be by-passed so the system can be tested for gravity circulation operation. The influence of the circulation ratio on the heat transfer and the pressure drop can be first measured and analyzed over the medium temperature simulator where it will be easier to evaluate the effect on heat transfer by measuring the brine temperatures. The obtained results from the simulator will be used as guidelines to test the proper range of circulation ratios on the display cabinets. The figure below shows the test circuit.

5 kW5 kW

Figure 24 Schematic diagram of the circulation rate test circuit

32


The available height in the laboratory is limited to 4 m and the ammonia unit is fixed on top of the CO2 unit due to the thermosyphon loop on the CO2 side of the cascade condenser. The ammonia unit also requires a certain head for its thermosyphon on the cascade condenser side. This means that there will be limited head for the CO2 for gravity circulation, about 1.7 m, which might not be sufficient to test the system for gravity circulation at full capacity and a compromise for smaller capacities might be needed. The medium temperature simulator is used to test for different circulation rates. The fixed load provided in the simulator makes it possible to test of a wide range of circulation ratios. Carioles mass flow meter is used to measure the mass flow of refrigerant, the flow rate at low circulation ratios of about 2 is very small compared to the measuring range of the mass flow meter. Due to the continuous generation of bubbles in the CO2 line it was not possible to perform the zero adjustment reading on the flow meter. For full capacity of the simulator the mass flow that is required to result in a circulation rate of 1 is 0.026 kg/s; less than 5% of the range of the mass flow meter where the estimated error is expected to be high. The valve located before the medium temperature simulator is used to control the mass flowing through the simulator. In order to reach the point of circulation 1 the valve was almost closed so that superheating is achieved in the simulator, it can be seen in the sight glass that there is no two phase flow in the circuit. Then the valve has been opened on small steps until the superheat was eliminated and the sight glass shows no liquid. It was noticed that slight opening of the valve after this point was reached resulted in liquid drops visual in the sight glass. The value of the mass flow reading at that point was used as the reference value to calculate the circulation ratio. The mass flow showed a reading of about 0.022 kg/s. Circulation rate have been changed between one and slightly over 14, the reason that this high value was reached was that the valve was opened gradually to identify at which point the heat transfer starts to change, the valve was fully open at the maximum circulation rate. Figure 25 shows the pressure drop across the simulator at different steps of circulation rate. As can be seen from the figure, the waiting time for each step is rather short, around 20 minutes, especially for the steps with low circulation rates. In order to confirm the accuracy in the data generated in this test another test have been run where the run time was longer, at least 1 hour, for less steps than the first test, Figure 26.

33


Pressure drop and CR

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0,60

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1,00

1,20

1,40

1,60

8:48:29 9:24:29 10:00:29 10:36:29 11:12:29 11:48:29 12:24:29 13:00:29 13:36:29 14:12:29 14:48:29 15:24:29 16:00:29

dP (b

ar)

1

2

3

4

5

6

7

8

9

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11

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13

14

15

16

CR

dP

CR

Figure 25 Simulator’s pressure drop and circulation ratio for short waiting time test

Pressure drop

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9

11

13

15

10:19:1210:55:1211:31:1212:07:1212:43:1213:19:1213:55:1214:31:1215:07:1215:43:1216:19:1216:55:12

dP (b

ar)

0,00

0,20

0,40

0,60

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1,00

1,20

1,40

CR

CRdP

34


Figure 26 Simulator’s pressure drop and circulation ratio for long waiting time test As can be seen from the figures the pressure drop across the heat exchanger increases with increasing the mass flow of refrigerant. Figure 27 relates the pressure drop as a function of CR, the plots in the diagram shows agreement between short and long run time tests.

Pressure drop vs CR

0,00

0,20

0,40

0,60

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1,00

1,20

1,40

1,60

0,00 1,00 2,00 3,00 4,00 5,00 6,00 7,00 8,00 9,00 10,00 11,00 12,00 13,00

CR

dP (b

ar)

Short Test PeriodLong test Period

Figure 27 Simulator’s pressure drop at different circulation ratio for short and long waiting time

Running at high circulation rate and having a high pressure drop could be justified if the heat transfer in the evaporator would improve with high circulation rate where an optimum circulation rate may exist. Measuring the brine temperatures around the simulator it was possible to evaluate the influence of the circulation rate on the heat transfer in the heat exchanger. Figure 28 shows the brine temperatures around the heat exchanger, the evaporation temperature of CO2 and the logarithmic mean temperature difference for the test with short waiting time test.

35


-11

-9

-7

-5

-3

-1

1

3

5

7:12:00 8:24:00 9:36:00 10:48:00 12:00:00 13:12:00 14:24:00 15:36:00 16:48:00

Tem

pera

ture

(C)

LMTDT Brine InT Brine OutT CO2

Figure 28 Temperatures around simulator

As can be seen from the figure, within the tested range the circulation ratio has almost no influence on the heat transfer in the heat exchanger; LMTD had almost constant value of about 3.6ºC along the test. Plot of the LMTD versus the circulation ratio for the two tests is presented in Figure 29 where the influence of the circulation ratio can be seen in a clearer way, the influence is very small where the LMTD caries within a range of 0.2ºC which is insignificant.

36


LMTD vs CR

3,3

3,4

3,5

3,6

3,7

3,8

3,9

4

0,00 1,00 2,00 3,00 4,00 5,00 6,00 7,00 8,00 9,00 10,00 11,00 12,00 13,00

CR

LMTD

C

Short Test PeriodLong Test Period

Figure 29 Simulator’s LMTD at different CR for long and short waiting time tests

As can be seem from the results above increasing the mass flow of refrigerant had an insignificant improvement on heat transfer. Increasing the circulation ratio resulted in an increase in the pressure drop across the heat exchanger without improvement on the heat transfer; this indicates that the circulation ratio should be chosen as low as possible to ensure complete evaporation at the highest load expected. There was no optimum operating circulation ratio found similar to the case of conventional brines in secondary systems, this gives more flexibility in choosing the operating conditions for CO2 secondary system, especially that its pressure drop is much lower than it is for brines. Similar test for low circulation rates will be conducted on the medium temperature level display cabinets. The pressure drop will be evaluated and it will be easier to compare it to theoretical calculation since the geometry in the cabinets’ tubes is much easier to apply in the theoretical model. The heat transfer is not expected to change in a measurable way in the cabinets and this will be observed from any change in air temperatures around the cabinet’s evaporator. 9.2 Cascade Condenser On the cascade condenser there are three main arrangements to be tested concerning the way CO2 condenses. The first arrangement is the one in Figure 30, where the hot gas return from the low stage is passed directly through the cascade condenser after mixing with the saturated vapor from the CO2 tank. A line has been introduced, shown in the figure as red dashed line, where some of the pumped liquid in the medium temperature level can be flashed into the hot

37


gas line before entering the cascade condenser so it will be de-superheated and that may result in more favorable temperature profile in the heat exchanger.

Figure 30 Schematic of a cascade condenser arrangement

The other arrangement is where CO2 is condensing in a thermosyphon loop, Figure 31a, where the two phase CO2 return line ends in the accumulation tank and the hot gas return from the low stage passes through the liquid of the tank so the hot gas will be de-superheated by boiling off some of the tank’s liquid. The third arrangement, Figure 31b, is where the return line from the medium temperature level is passing directly through the cascade condenser after mixing with the hot discharge from the low stage.

(a) (b)

Figure 31 Schematic diagrams of two cascade condensation arrangements The system has been all the time running on the thermosyphon arrangement described in Figure 31a. Due to the resulted high discharge temperature of CO2 the two other arrangements have not been tested yet. It will be possible to perform the tests with better CO2 compressor that has higher isentropic efficiency. The temperature at the compressor inlet should be reduced and this can be achieved by running the freezers at the minimum superheat setting

38


where the system will have stable operation, also the external superheat which is due to heat sink from the environment should be reduced, in average about 7°C was observed. Figure 32 shows the temperature values across the cascade condenser; the difference between the “hot” CO2 and the “cold” ammonia is slightly higher than 2°C. In this test the load was only from the low temperature side and the medium temperature was off.

Cascade Condenser Temperatures

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75,00 175,00 275,00 375,00 475,00 575,00 675,00

Tem

pera

ture

(C)

Temperature differenceCO2Ammonia

Time span 09:10:34-15:23:26

Figure 32 Temperatures across the cascade condenser

In the applications where CO2 is used in indirect circuit the two arrangements that have been used are shown in the schematics below. The arrangement A in the figure below has been used in the early stages of applying CO2 in indirect solutions for freezing temperature applications and then has been abandoned in favor of the natural circulation solution/thermosyphon.

Figure 33 Basic schematic of two possible arrangements for the CO2 indirect circuit

39


The arrangement in Figure 33b has been tested and the results are similar to the ones plotted in Figure 32. The influence on heat transfer in the cascade condenser by using the other indirect system arrangement will be tested by applying the system arrangement in Figure 34. The performance of the cascade condenser will be tested for different loads on both arrangements which will change the mass flow of refrigerants on both sides of the cascade condenser and the effect on heat transfer will be analyzed in a similar way in the Figure 32. The stability of operation which has been one of the reasons for favoring the thermosyphon arrangement will be investigated.

Figure 34 Basic schematic of forced condensation arrangement on the cascade condenser

The temperature drop that is measured and observed while operating the cascade condenser indicates a good heat transfer that yields a low temperature difference across the cascade condenser. This is mainly due to the favorable conditions for exchanging heat on the sides of the heat exchanger. CO2 enters the cascade condenser as saturated vapor and ammonia boils off along the heat exchanger from saturated conditions at the inlet. Further evaluation at different loads will be performed and other arrangements available in the system will be tested. 9.3 Flashing Gas in Liquid Lines Long supply lines to the freezing cabinets are usually the case in this application which means that there will be a pressure drop and some of the refrigerant might flash before the expansion valve resulting in unstable operation. An internal heat exchanger is usually used to tackle this problem by adding some sub-cooling before the distribution lines, Figure 35. Another possibility is to add some head to the liquid and this is achieved by connecting the liquid supply to the expansion valves after the medium temperature circulation pump, Figure 36, in this case the additional pumping power accompanying this solution will be very small since the CO2 pump is larger than needed. Both solutions are planned to be tested.

40


Figure 35 Basic schematic of internal heat exchanger solution

Figure 36 Basic schematic of the solution where head is added to the liquid before expansion valve So far the system has been running without pump heat or internal heat exchanger. The pressure drop and heat leaking in the liquid supply line proved to be small and the available head in the tank was enough to overcome it. Both arrangements in the figure above have been tested and they work properly, when running with internal heat exchanger the discharge gas temperature was noticed to be very high. This has to do with the low isentropic efficiency of the CO2 compressor. 9.4 Freezing Cabinets Control The two freezers in the system are identical and equipped with electronic expansion valves. The freezers were delivered with a control system that measures the evaporation temperature on the surface of the evaporator tube. In this case the thermocouple is placed at the first bend after the expansion valve. Figure 37 is a schematic diagram of the freezer evaporator where the positions of the measuring points are indicated.

41


t2

t_sh

Figure 37 Basic schematic of the freezer’s evaporator with the measuring points indicated

The freezers showed occasional unstable behavior, especially at start up, while running the system with this control method, this was more likely to happen at low superheat set point. While running the system at the lowest set point of 3°C the air temperature in the freezers started to increase up to a certain value and then decrease again and the behavior is repeated in a cyclic manner. This is plotted in Figure 38 where it can be seen that the air temperature difference is decreasing which indicates loss in cooling capacity.

-40

-30

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-10

0

10

20

13:48:00 14:16:48 14:45:36 15:14:24 15:43:12 16:12:00

Tem

pera

ture

(C)

Superheat Air in Air out Freezer surface-controller Freeze

Superheat=3°

Figure 38 Temperatures around the freezer cabinet with temperature based controller. Superheat set value is 3ºC.

42


The controller starts controlling the system for the highest superheat set value, usually set between 10-15°C, and then the valve opening is adjusted targeting the lowest superheat set value, where the system is usually operating at in a stable condition. When the system reaches the lowest superheat set point this behavior is triggered. This can be seen in the superheat value in the figure above where the controller is trying to control the system to operate at the superheat value of 3°C. During the unstable operation there was not enough refrigerant flowing in the evaporator which led to the refrigerant being superheated at the point where the evaporation temperature should be measured. The evaporating temperature measured by the pressure transducer is plotted in the diagram which indicates the difference between the measurement of the real evaporating temperature from the pressure reading and the one that is measured on the surface of the heat exchanger tube. The superheat set point of 3°C is very low compared to practice but the same behavior was repeated with higher superheat set values but with fewer oscillations after the start up period. A sample of the tests is shown in Figure 39 where the superheat set value of 7°C was used. The unstable behavior occurred only once at start up of the test.

-40

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0

10

11:24:00 11:52:48 12:21:36 12:50:24 13:19:12 13:48:00 14:16:48 14:45:36 15:14:24

Tem

pera

ture

(C)

T SH SurfaceT Air InT Air outT Freeze SurfaceT Freeze Sat Press

Superheat=7

Figure 39 Temperatures around the freezer cabinet with temperature based controller. Superheat set value is 7ºC. Another method of measuring the evaporation temperature is by measuring the evaporation pressure and this will be the input signal to the expansion valve controller. A schematic diagram of the pressure based controller method is presented in Figure 40. In this case the

43


actual evaporation temperature is measured instead of the surface temperature which is about 2°C higher than the real value; this can be seen in the plots in the two figures above.

Figure 40 Basic schematic of the freezer’s evaporator with the measuring points indicated

The controller was changed on one display cabinet while the other was kept on the surface temperature measurement method. The freezers were run at the same conditions and at the same set point. The influence of the superheat value is tested, a value of 8°C was used and the two cabinets were running with good stability as can be seen in Figure 41 which shows the air and brine temperatures for the cabinets and the simulator. The simulator used the pressure based control all the time.

Freezers and Simulator Temperature

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Tem

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ture

(C)

Air in Press ControlAir out Press ControlAir in Temp ControlAir out Temp ControlBrine in Press ControlBrine out Press Control

Superheat=8-10°C

Time span 09:10:34-11:12:43

44


Figure 41 Temperatures around the freezers and simulator. Superheat set value is 8ºC. Figure 42 and Figure 43 shows evaporating temperatures, superheat values and the superheat temperature measured at the exit of the evaporator for the freezers and the simulator. The variation in the superheat temperature in the freezer with surface temperature control, in Figure 42, may be due to the fact that the surface temperature controller senses lower superheat than the pressure one and starts regulate at earlier time than the pressure based controller. As can be seen in Figure 42, the measured superheat temperature at the exit of the evaporator in case of the simulator is fluctuating at start up which may be due to the small size of the heat exchanger where the response to the expansion valve opening or closing is faster and more prominent than the case of the long freezer heat exchanger.

Freezers Controlers Paramters

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Tem

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SuperHeat Press ControlSuperHeat Temp ControlEvap exit Press ControlEvap exit Temp ControlFreezing Temp ControlFrezer Press Control

Superheat=8-10°C

Time span 09:10:34-11:12:43

Figure 42 Temperature values input to the controller in the freezers. Superheat set value is 8ºC.

45


Simulator Controlers Paramters

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-20,00

-10,00

0,00

10,00

20,00

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40,00

0,00 100,00 200,00 300,00 400,00 500,00 600,00 700,00

Tem

pera

ture

(C)

Super HeatEvap exitFrezering

Superheat=8-10°C

Time span 09:10:34-11:12:43

Figure 43 Temperature values input to the controller in the simulator. Superheat set value is 8ºC.

When the superheat temperature is reduced to 5°C in an attempt to trigger the unstable behavior both freezers started to show cyclic increase and decrease in the air temperature, Figure 44 and Figure 45 shows the air and superheat temperatures on both cabinets. In case of surface temperature control the oscillations had higher peaks and the air exit temperature was close to 0°C, while in the other case the value did not go over -12°C. And the superheat in the second case is oscillating between 2 and 10°C while in the first case the superheat value goes below zero and then increase up to 20°C. The superheat high set value for the controller is 10°C in both cases.

46


Freezers Temperature and Superheat

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11:45:36 12:14:24 12:43:12 13:12:00 13:40:48 14:09:36 14:38:24 15:07:12

Tem

pera

ture

(C) Superheat

Air inAir out

Superheat=5-10°C

Figure 44 Temperatures around the freezer with pressure based controller. Superheat set value is 5ºC.

Freezers Temperature and Superheat

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11:45:36 12:14:24 12:43:12 13:12:00 13:40:48 14:09:36 14:38:24 15:07:12

Tem

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(C) Supeheat

Air inAir out

Superheat=5-10°C

Figure 45 Temperatures around the freezer with temperature based controller. Superheat set value is 5ºC.

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The parameters that the controllers read in both cases are presented in Figure 46 and Figure 47, as can be seen in Figure 46 the evaporating temperature read is increasing due to refrigerant superheat and the controller reads the wrong evaporation temperature and fails to control properly.

Freezers Controllers Parameters

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Tem

pera

ture

(C)

Super HeatFreezerEvap Exit

Superheat=5-10°C

Figure 46 Temperature values input to the pressure based controller in the freezer. Superheat set value is 5ºC.

Freezers Controllers Parameters

-30

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Super HeatFreezer Temp ControlEvap Exit

Superheat=5-10°C

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Figure 47 Temperature values input to the temperature based controller in the freezer. Superheat set value is 5ºC. In the case of pressure based controller it measures a higher superheat value than the actual one, which can be seen in the figures above as a difference between surface measured evaporation temperature and the pressure related one. This is true assuming that the same temperature difference exists at the superheat measuring point. This gives the pressure based controller higher margin of stability, on the other hand the temperature based controller may be closer to measuring the real superheat value but in the cases discussed above at some conditions it measures wrong evaporation temperature. Also the response to changes or fluctuations in evaporation temperature would be faster in case of the pressure based controller. In the case of the simulator the controller managed to regulate the superheat value adequately at the 5°C set value, this can be seen in Figure 48. This may be due to the fast response at the evaporator exit to any changes to the degree of opening of the expansion valve which consequently readjusts itself accordingly.

Freezer Simulator Brine Temperature and Superheat

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Tem

pera

ture

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SuperheatBrine inBrine out

Superheat=5-10°C

Figure 48 Temperatures around the simulator with superheat set value of 5ºC.

The measurements and experiences in running the display cabinets with the two controllers indicate a more favorable conditions in case of the pressure based controller. The controller successfully kept the superheat value within the set range, it had lower stability at low superheat set values but still not as unstable as the temperature based controller. Measuring the evaporation temperature on the surface of the tube of the evaporator proved to be the source of instability in the controller; the point measured superheated vapor when the mass flow was small in the heat exchanger.

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9.5 Safety Tests The main two safety concerns when dealing with CO2 in such an application is the problem of the rising pressure at standstill and CO2 concentration level during and after a leakage accident. Several valves at different points of the system where CO2 exists as gas, liquid, and two-phase are installed. The valves will be opened in order to simulate leakage of a pipe breaking. The rate of leakage will be observed together with the rate of increase in concentration at different levels in the room. The concentration will be measured at different zones in the space to try to identify the severity of an accident and the best place to install the CO2 indicators for the alarm system. The formation of dry ice at the leakage point may block the leakage or reduce its rate. The pressure in the system will be allowed to rise to levels where the bleed valves will be opened and then the time for the system to reach this state will be observed and the concentration rise in the room will be recorded. Also, the frequency of this operation and the amount of lost charge will be reported. 9.6 Flooded Versus Direct Expansion Evaporation The performance analysis of the medium temperature flooded evaporators will be evaluated against the direct expansion freezers, accordingly, it will be possible to decide about how a flooded freezer may perform and estimate the improvements that may be gained in each solution. In the preliminary design it is made possible to transfer the evaporation in the freezers from direct expansion to flooded type by adding a low pressure tank, a pump and by-passing the expansion valves.

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10. THEORETICAL MODEL Along with the experimental rig, a mathematical model has been developed to simulate the experimental system. At the design stage of the system, the theoretical model facilitated design calculations concerning sizing of components and it will be evaluated against the experimental results when the system will be running. If the theoretical model agrees with the experimental results, this will give confidence in evaluating the performance of the system by varying certain parameters and will help in evaluating design modifications. 10.1 Model Description The model is written in Engineering Equation Solver (EES) and uses the manufacturers’ data to calculate the performance of each component. The boundary conditions and the cooling capacities of the system are input variables to the model which are inserted in the performance equations and efficiency curves to calculate energy consumption, capacities, and efficiencies of different components and the system as a whole. 10.2 Overview of Calculations and Assumptions Two phase flow pressure drop calculations use Friedel’s correlation which is based on the two-phase multiplier. The Clapeyron equation is used to convert the pressure drop to an equivalent change in saturation temperature. Single-phase flow pressure drop is calculated using the Gnielinsky correlation for the friction factor in turbulent flow. More details about the pressure and temperature drop calculations for CO2 can be found in (Sawalha and Palm 2003). The heat exchangers’ performances have been simulated by calculating the effectiveness based on the supplied manufacturer’s performance data for the heat transfer coefficient and the heat transfer area (Arias 2005). These properties are assumed constant and the evaporating and condensation temperature along the heat exchangers are also assumed constant. In the cascade condenser and in the thermosyphon mode at the CO2 side the logarithmic mean temperature difference method has been calculated since the temperature is constant at both sides of the heat exchanger. The compressors are presently assumed to have constant isentropic and volumetric efficiencies but in a later stage the efficiencies will be introduced in the model based on the performance curves for the compressors. In the display cabinets the flow is assumed to be equally distributed in the cabinets’ loops. The pressure drop in the medium temperature display cabinets is calculated assuming single phase flow (liquid) in the first half of the loop’s pipe and two phase flow with the exit quality in the other half, consequently the total pressure drop is the sum of the two pressure drops. In the deep freeze cabinets, where dry expansion takes place, the pressure drop is calculated assuming two phase flow in the first half of the pipe with the inlet quality to the evaporator and in the second half it is assumed to be single phase flow with the exit conditions of superheated vapor, the total pressure drop is the sum of the two values. These assumptions are made to simplify the calculations and are only used for the estimation of the pressure drops.

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10.3 Design Calculations Pressure and temperature drop calculations have been used to size the distribution lines of the system. Figure 49a and Figure 49b show results for the medium circuit calculation with a circulation ratio of 3 and the length of each line being 20 m. The figures show plots of the pressure and temperature drops for the liquid supply and the two-phase return lines using different pipe diameters. It is possible to change the circulation ratio and obtain such plots for each value; the results for the value of 3 have been selected because it fits almost in the middle of the range of the values for the circulation ratio to be tested. The plotted total pressure drop includes the losses in the fittings along the line, which were calculated using the equivalent pressure drop coefficient method (Stoecker 1998).

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C]

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(a) (b)

Figure 49 Pressure and temperature drops for different tube sizes in medium temperature liquid supply (a) and return (b) lines

Similar calculations have been made for the low stage and a sample of the results for the single phase vapor in the suction line is presented in Figure 50.

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Figure 50 Pressure and temperature drops in the suction line for different tube diameters

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As can be seen from the plots, the pressure and temperature drops for CO2 are very small, even with small pipe sizes. The sizes of the major lines in the rig are 5/8” (15.8 mm) for the liquid supply line of the medium temperature stage, the return line has 1 1/8” (28.6 mm), the liquid line of the low stage is 3/8” (9.5 mm) and the suction line has a size of 5/8”. Based on these sizes it has been estimated that gravity circulation will be possible to achieve with the available head of 1.7 m. At standstill the system will be absorbing heat from the warm surroundings. The main volume of the system is assumed to be the tank since at standstill the charge will be accumulating in the tank. The heat flow from the 25ºC room into the -8ºC tank with different insulation thicknesses covering the tank is plotted in Figure 51. Assuming 50% of the tank filled with liquid and using 50mm of insulation the time for the system to reach the pressure of 35 bars, which is the limit that will trigger the bleed valves, is calculated to be about 12 hours. At standstill conditions the CO2 in the tank is cooled down by using cold brine circuit available in the laboratory. By using heat exchanger with a capacity of 1 kW the time required to remove the heat added from the ambient to the tank since standstill started is estimated to be about 50 min. In the calculations, the temperature of the tank is assumed to be constant during the heat transfer into the system, therefore in real conditions the time of pressure rise will be longer than calculated.

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winsulation [millimeter]

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]

Figure 51 Heat sink into the accumulation tank with different insulation thickness covering the tank

10.4 Some Simulation Results At this stage of the development of the model with the available manufacturing data some performance calculations can be obtained to estimate the overall performance of the system. The COP of the low stage is about 4, for the ammonia unit the COP is 2.6 and the total COP is 2.1. The total COP is calculated as the ratio of the total cooling capacity and the total power consumption. With a circulation ratio of 2, the pump power consumption is calculated to be 12 W with 50% of assumed pump efficiency.

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11. CONCLUSIONS AND FUTURE WORK The chosen NH3/CO2 system for supermarket refrigeration is being built in a laboratory environment which can simulate the conditions in a real installation. Different variations in the system have been implemented in order to compare different options and find an optimum system design for the usage of CO2 in the low and medium temperature stages of a supermarket refrigeration system. Several tests have been run for overall evaluation of the system and for the main components. A mathematical model has been developed to simulate the performance of the system and facilitate detailed system design calculations. The experimental and computational models will be adjusted while running the system in order to bring the two models closer to each other to be more comparable. The closer the theoretical model will get to simulate the real case the more confidence it will give to use it for adjustments in the system for the optimum operation. The energy balance in the system has been verified and the efficiencies of the compressors have been measured and calculated. The obtained efficiencies of the compressors will be used to calculate cooling capacities and shaft power consumption of the ammonia compressor. Due to unfavorable operating environment the CO2 compressor had to run in its volumetric and isentropic efficiencies have been measured to be much less than expected. System’s high stage, low stage and total COP’s are measured and calculated. The measured COP’s of the system are lower than expected and this is mainly due to running the system at partial load which is companioned with some mechanical and electric motor losses; at loads close to full capacity ammonia unit and total system COP is expected to improve. With CO2 compressor that has the design efficiencies low and total COP’s are expected to improve. The circulation ratio in the simulator has been changed and its influence on heat transfer and pressure drop has been evaluated. Increasing the mass flow of refrigerant to have circulation ratios higher than one had insignificant improvement on heat transfer. The pressure drop across the heat exchanger increased and no optimum operating point have been found. Therefore, the circulation ratio should be chosen as low as possible to ensure complete evaporation at the highest load expected. Similar test for low circulation rates will be conducted on the medium temperature level display cabinets, mainly to evaluate the pressure drop in its tubes. The temperature drop that is measured and observed while operating the cascade condenser in the thermosyphon mode indicates good heat transfer conditions. This is mainly due to the favorable conditions for exchanging heat on the sides of the heat exchanger; CO2 enters the cascade condenser as saturated vapor and ammonia boils off along the heat exchanger from saturated conditions at the inlet. Further evaluation at different loads will be performed and other arrangements available in the system will be tested. The measurements and experiences in running the display cabinets with the two controllers indicated more favorable conditions in case of the pressure based controller. The controller successfully kept the superheat value within the set range, it had lower stability at low superheat set values but still not as unstable as the temperature based controller.

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Safety analysis of the system with different scenarios will be performed both theoretically and practically and will provide more detailed information about the behavior of CO2 and allow development of methods for safe handling. Based on the results of the tests performed up to this stage, weak and strong points of the system have been identified where the system will be modified for the best of its performance at the selected basic arrangement. The analysis to follow will be focusing on testing the possible arrangements available in the test rig and to evaluate parametric analysis of the main system’s components.

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12. REFERENCES ASHRAE 1997. ASHRAE Handbook, Fundamentals. Sawalha, S. and Palm, B. 2003, Energy Consumption Evaluation of Indirect Systems With CO2 as Secondary Refrigerant in Supermarket Refrigeration, Proceedings of the 21st IIR International Congress of Refrigeration, Washington, D.C., USA

Pierre, B.: ”Kylteknik, allmän kurs”, Inst. Mekanisk värmeteori och kylteknik, KTH, Stockholm, 1982 (in Swedish)

Granryd, E., Ekroth, I., Ludqvist, P., Melinder, Å., Palm, B., and Rohlin, P. 2003, Refrigeration Engineering, Department of Energy Technology, KTH, Stockholm, Sweden.

Arias, J. 2005. Energy Usage in Supermarkets Modelling and Field Measurements, Doctoral thesis, Department of Energy Technology, Royal Institute of Technology. Stockholm, Sweden. Stoecker, W. F. 1998. Industrial Refrigeration Handbook, McGraw-Hill. Zhao, Y., M. Molki, et al. 2000. Flow Boiling of CO2 in Micro channels. ASHRAE Transactions 106(1): 437-445.

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project 2: co in supermarket refrigeration - kth · the project “co2 in supermarket...

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