Embed Size (px)
Citation preview
Royal Institute of Technology (KTH)
Investigations of Heat Recovery in Different Refrigeration System Solutions in Supermarkets Effsys2 project final report
Samer Sawalha and Yang Chen 2010‐06‐30
2
Table of Contents 1 Introduction ........................................................................................................................ 4
1.1 Background .................................................................................................................. 4
1.2 Objectives .................................................................................................................... 4
1.3 Methodology ................................................................................................................ 5
1.4 Project partners and acknowledgment ......................................................................... 5
2 Description of refrigeration and heat recovery systems ..................................................... 6
2.1 Energy use in supermarkets and typical heating requirements .................................... 6
2.2 Refrigeration system solutions .................................................................................... 8
2.2.1 Conventional R404A system and CO2 pump circulation .................................... 8
2.2.2 Cascade systems with CO2 .................................................................................. 9
2.2.3 CO2 trans-critical systems ................................................................................. 11
2.3 Heat Recovery System Solutions .............................................................................. 12
2.3.1 Fixed head pressure (FHP) ................................................................................. 13
2.3.2 De-superheater (DSH) ........................................................................................ 13
2.3.3 Heat pump cascade (HPC) ................................................................................. 14
2.3.4 Heat pump cascade for sub-cooling (HPSC) ...................................................... 14
3 Field measurements .......................................................................................................... 16
3.1 Calculations of heat recovery .................................................................................... 16
3.2 CO2 trans-critical system 1 ........................................................................................ 16
3.2.1 System design and operation .............................................................................. 16
3.2.2 Existing mode of operation and selection of operating conditions for heat recovery with heat pump .................................................................................................. 17
3.2.3 Recoverable heat with heat pump solution ........................................................ 20
4.2 CO2 trans-critical system 2 ........................................................................................ 21
3.2.1 System design and operation .............................................................................. 21
3.2.2 Total heat recovery ............................................................................................. 23
3.2.3 Performance analysis of medium temperature unit (KA3) ................................ 25
4.2.4 Performance of booster unit (KAFA1) ............................................................... 27
3.3 Conclusions ............................................................................................................... 28
4 Computer simulation modelling ....................................................................................... 30
4.1 System performance in floating condensing-No heat recovery ................................. 31
4.2 System performance in fixed head pressure (FHP) heat recovery ............................ 33
4.3 System performance in heat pump cascade (HPC) heat recovery ............................. 34
3
4.4 System performance in heat pump cascade for sub-cooling (HPSC) heat recovery . 36
4.5 System performance in de-superheater (DSH) heat recovery ................................... 37
4.6 Annual energy consumption calculations .................................................................. 46
4.7 Conclusions ............................................................................................................... 48
5 Experimental measurements of CO2 heat pump test rig .................................................. 48
5.1 System Layout ........................................................................................................... 50
5.2 Working principle ...................................................................................................... 50
5.3 Methodology .............................................................................................................. 51
5.4 Testing results ............................................................................................................ 54
5.4.1 Operating conditions .......................................................................................... 54
5.4.2 Overall system performance ............................................................................... 54
5.4.3 Results for the compressor part .......................................................................... 56
5.4.4 UA value with the flow rate ............................................................................... 64
5.5 Comparison with other studies .................................................................................. 65
5.5.1 Case study 1 ....................................................................................................... 66
5.5.2 Case Study 2 ....................................................................................................... 66
5.5.3 Case Study 3 ....................................................................................................... 68
5.6 Conclusions ............................................................................................................... 69
6 Discussions and conclusions ............................................................................................ 70
7 References ........................................................................................................................ 71
4
1 Introduction
1.1 Background
Supermarkets are intensive energy consumers with constantly increasing number of installations. About 50% of the energy consumption in the supermarket is absorbed by the refrigeration system to cover the cooling demands, simultaneously; heating is needed in the supermarket where the rejected heat from the refrigeration system is usually higher than the needs. It is an interesting possibility to utilize the rejected heat from the refrigeration system to cover the heating needs in supermarkets.
The environmentally friendly natural refrigerants are seen to be a long term solution in many refrigeration applications as well as in supermarkets. The new solutions provide interesting potential for reclaiming heat from the refrigeration system. Solutions such as parallel, centralized, cascade, direct, indirect, single‐, two‐, or multiple‐stage exist in supermarkets today where each solution will have different temperature levels at which heat is rejected. It will also have different capacities at different temperature levels depending on the design of the plant, refrigeration loads, ambient conditions, and the type of compressor.
Previous research conducted at the Energy Technology Department at KTH shows that when comparing cooling efficiencies in supermarkets, solutions which are based on natural refrigerants have good potential in energy savings compared to conventional systems. Additionally, certain systems, such as trans‐critical CO2, have good heat rejection characteristics with relatively high efficiency when used in certain heat pump mode. However, the system is sensitive to the conditions/requirements under which it operates. Therefore, it is important to study its performance under the supermarket heating requirements.
This project investigates the heat recovery performance of the new system solutions that are mainly based on natural refrigerants in supermarket refrigeration. The new system solutions with natural refrigerants provide fresh possibilities for heat recovery due to the different temperature levels at which certain heat capacities are rejected. Thus, the investigated system modifications and optimizations will take into account not only the cooling efficiency but also the efficient heat recovery according to the needs in the supermarket.
1.2 Objectives The objective of this project is to investigate the heat recovery performance of the new refrigeration system solutions in supermarket applications. The focus is on environmentally friendly systems with natural working fluids, mainly CO2 trans‐critical systems. The project analyzes the temperature levels and capacities of rejected heat from different system solutions and investigates its matching with the heating needs in supermarkets.
Using simulation tools this project also aims at defining the system solution/s which has good energy efficiency for simultaneous cooling and heat recovery.
In order to verify the theoretical analysis experimental/field tests are planned for selected systems.
5
1.3 Methodology The work in this project is conducted based on the previous knowledge in refrigeration engineering developed at the Energy Technology Department in projects dealing with energy efficiency and alternative refrigeration system solutions in supermarket as well as heat pump applications. Experience in developing computer simulation tools and the development of a state‐of‐the‐art experimental test rig are important guide and resource in conducting the system analysis. Field measurements of selected systems are used to evaluate its performance and to have input parameters to the simulation models.
The work in this project is divided in three main work packages:
Field measurements
Two supermarkets operating with CO2 trans‐critical refrigeration systems have been selected for the analysis. The systems have different heat recovery methods. The evaluation of the systems focused on analyzing the performance of the system in both refrigeration and heat recovery modes. Computer simulation modelling has been used in validating the results of the field measurements and to evaluate the system performance under different cooling and heating requirements, which would not be possible to change/control in real installations.
Computer simulation modelling
Computer models have been developed to simulate the performance of each system solution under investigation. The refrigeration system solutions include, CO2 trans‐critical (parallel and booster), NH3/CO2 cascade, R404A/CO2 cascade, and R404A conventional. Most of the assumptions for the computer simulation models have been extracted from the field measurements data and the simulation models then has been used to evaluate the performance of the selected systems under different operating conditions.
Experimental investigations
A test‐rig has been built up at KTH, in the laboratory of Applied Thermodynamics and Refrigeration division, the test rig is used in the project to validate computer simulations and it is also used to evaluate various components such as heat exchangers, compressors, expansion valves and control systems for CO2 system. Different system configurations are evaluated to optimally match the heat rejection of the refrigeration system to the various possible heat sinks for best possible efficiency.
1.4 Project partners and acknowledgment The project has been conducted by the Energy Technology Department at KTH in cooperation with the following companies: AlfaLaval, Ahlsell, Nibe, IVT, SRM, Danfoss, Green and Cool, RANOTOR, Climacheck, ICA Sverige AB, and Huurre Sweden AB. The project is co‐financed by the Swedish Energy Agency (STEM).
Through the project close cooperation has been maintained with IUC Sveriges Energi‐ & Kylcentrum AB (IUC‐SEK).
I would like to thank all the project partners for their contributions to this work and Jaime Arias at KTH who helped in providing some data.
6
2 Description of refrigeration and heat recovery systems
2.1 Energy use in supermarkets and typical heating requirements In Sweden, approximately 3% of the national electricity consumed is used in supermarkets (Sjöberg, 1997). A typical supermarket in Sweden uses between 35‐50% of its total electricity consumption for refrigeration equipment (Lundqvist, 2000). Figure 2‐1 shows the energy use in a typical supermarket in Sweden, as can be seen in the figure, considerable portion of energy is consumed in refrigeration, lighting and HVAC, while refrigeration system rejects considerable amount of heat. The recovery of heat from the refrigeration system presents potential to reduce or recover heating needs for HVAC and service water.
Figure 2‐1: Typical energy use in a supermarket in Sweden (Arias, 2005)
Common heating applications in a supermarket consist of floor heating, heater for HVAC system, service water heater and in some applications, defrosting of evaporator coils. Some heating applications and temperature requirements are listed in the following table.
Table 2‐1: Temperature requirements in heating applications (Wulfinghoff D. R, 1999)
Applications Temperature level (oC)
Floor heating About 27
HVAC dimensioning temp 33
Service hot water 54
Hydronic system 70‐40
Theoretically, in refrigeration system heat can be recovered from compressor oil, de‐superheater and condenser. Potential of heat recovery depends on the quality and quantity of heat required and
Energy usage in a supermarket
Refrigeration, 47%
Lighting, 27%
HVAC, 13%
Kitchen, 3%
Outdoors, 5%
Others, 5%
7
experience varies on the percentage of high temperature that can be recovered. In an industrial ammonia vapour compression system only 11% of the total system heat rejection is available in superheat region and the remaining is rejected at lower quality (Reindl D. T & Jekel T. B, 2007). Minea V. suggested that up to 30% heat can be recovered at the superheat region (Minea V, 2007). Therefore the percentage of heat can be recovered depends on the refrigerants and the system solution.
Heat recovery system solutions in supermarket are used mainly to heat the space air. The practical experience indicated that though seemingly high quantity of heat is rejected by supermarket refrigeration systems, only 40‐70% of the necessary heat can be recovered (Arias, 2005). Arias also suggested that refrigeration system not operating continuously to be the likely reason for less recovery of heat. In a typical Swedish supermarket HVAC system and refrigeration system are installed and operated by different companies. Therefore HVAC and refrigeration systems are separated from each other by heat exchanger which has been cited as another reason for the low heat recovery percentage.
An important parameter in the analysis of the heat recovery potential for refrigeration system is the match between the heating and cooling demands. At low outdoor temperatures the relative humidity is low and so is the cooling capacity of the system, at the same time the heating demand in the supermarket increases. Therefore, the relative size/capacity of the refrigeration system to the size/heating needs in the supermarket is an essential parameter to investigate in order to achieve efficient simultaneous cooling and heating.
The regulations on synthetic refrigerants have forced major changes in the refrigeration, air conditioning, and heat pump industries. Generally, the new situation required the old refrigerants to be replaced, systems to be tighter and new system solutions which require less refrigerant charge to be introduced. Nevertheless, the energy consumption of the systems should be kept as low as possible.
HFC refrigerants were expected to be an acceptable replacement for the phased out CFC and HCFCs but they became subject to regulations in some countries due to their Global Warming Potential (GWP). In Sweden, R404A, which has GWP value of 3784, has been intensively used in supermarket installations as a replacement for the environmentally harmful refrigerants, mainly R22. In order to reduce the refrigerant charge, systems that use an indirect solution were applied. Systems’ tightness has been improved due to taxes enforced on leaking HFC refrigerants.
Natural refrigerants are seen as a potentially permanent solution where CO2 is the one that fits best in supermarket applications, mostly due to safety reasons, as it can be directly used in public areas. CO2 was first used as a secondary working fluid in indirect system solutions, the first CO2 indirect system plant was implemented in Sweden in 1995.
The knowledge and experience gained from early research work on CO2 and the early installations of CO2 in commercial applications promoted its wider application in supermarkets with different system solutions. Cascade systems with CO2 in the low stage and trans‐critical solutions where CO2 is the only working fluid have been applied in Sweden in many installations.
8
Different solutions have been applied in the CO2 cascade and trans‐critical concepts which resulted in an interesting wide variety of conventional and new refrigeration system solutions in Sweden. The different refrigeration system solutions will then have different performance in heat recovery mode, it also provide fresh possibilities for heat recovery which is different than the conventional techniques.
The following sections describe the main categories of the refrigeration system solutions in Swedish supermarkets and the main techniques for heat recovery.
2.2 Refrigeration system solutions In general, two temperature levels are required in supermarkets for chilled and frozen products. Product temperatures of around +3°C and –18°C are commonly maintained. In these applications, with a large difference between evaporating and condensing temperatures, the cascade or other two‐stage systems become favourable and are adaptable for the two‐temperature level requirements of the supermarket. The following sub‐sections describe the conventional and the main CO2‐based solutions in supermarkets.
2.2.1 Conventional R404A system and CO2 pump circulation Systems using R404A are referred to as conventional technologies where it is used as replacement for harmful synthetic refrigerants. The R404A system consists of two separate circuits; DX at the low temperature unit and with brine at the medium temperature level. A heat exchanger is connecting the medium and the low temperature stages of the system to further sub‐cool the liquid coming out of the low temperature condenser or sub‐cooler.
Due to the steepness of the isentropic compression lines for R404A two‐stage compression with inter‐cooling has very little influence on improving the COP of the medium and low temperature levels. Therefore, two‐stage compression is not a conventional solution with R404A.
The following figure is a simple schematic of the system.
Figure 2‐2: Schematic diagram of a conventional R404A system
Due to the regulations on the HFC charges in the refrigeration systems the use of indirect system at the low stage has been applied in order to avoid the use of R404A in DX solutions. CO2 has been used
9
in many installations as the secondary working fluid where the operating pressure is reasonable (11 bars at ‐37°C). Figure 2‐3 is a simple schematic of the R404A‐CO2 indirect refrigeration system.
Figure 2‐3: Schematic diagram of R404A system with CO2 pump circulation at the low temperature level
From gaining experience in operating CO2 at freezing temperature levels, the same concept has been applied to medium temperature levels, as reported by Madsen et al. (2003). The temperature required for chilled food is usually around +3°C and should not go higher than 7°C for long periods of time. Assuming 7K of temperature difference in the display cabinet results in CO2 operating at –4°C, which corresponds to around 31 bars. In this case, components that can withstand 40 bars (corresponding to saturation temperature of 6°C) can handle CO2 with acceptable safety margins. However, such system solution is not applied in Sweden.
Only the conventional system solution will be analyzed in this study because of the similarity on the heat rejection side between the two systems.
2.2.2 Cascade systems with CO2 Cascade systems with CO2 in the low‐temperature stage have been applied in several supermarket installations in Sweden. Two main possible arrangements are the following:
R404A‐CO2 cascade
In this system arrangement, which exists in several installations in Sweden, the refrigerant in the high‐temperature stage is R404A. The medium temperature circuit uses a conventional single phase secondary working fluids. CO2 is the working fluid in the low‐temperature circuit where it rejects the heat to the brine at the medium temperature level. The following plot is a simple schematic of such system.
10
Figure 2‐4: Schematic diagram of R404A‐CO2 cascade system with brine at the medium temperature level
NH3‐CO2 cascade
A solution which is based on natural refrigerants and an alternative candidate to conventional systems is the NH3/CO2 cascade concept. Figure 2‐5 is a schematic diagram of such a system solution which has been built and tested in a laboratory environment (Sawalha, 2008). In this system concept CO2 pressure levels are acceptable; when condensing at ‐3°C pressure is about 32 bars. At this temperature level, as can be seen in Figure 2‐5, the cooling load in the medium temperature level is provided by circulating CO2 that accumulates in the tank.
Figure 2‐5: Schematic diagram of NH3‐CO2 cascade system with CO2 as the refrigerant for the medium temperature level
11
This system has not been installed in supermarkets in Sweden yet but has been included in part of the analysis in this study in order to demonstrate its relative performance in heat recovery mode. However, it has been demonstrated that this system has very good performance in refrigeration mode, 40‐50% higher COP compared to conventional solutions (Sawalha, 2008).
2.2.3 CO2 transcritical systems Due to the widespread interest in CO2 as an alternative to synthetic refrigerants in the two major refrigerant consuming applications, i.e. mobile air conditioning and commercial refrigeration, components which are specially designed to handle CO2 have become increasingly available and competitive in price. This made it possible to build CO2 trans‐critical systems for supermarkets, the main two arrangements applied in Swedish supermarkets are:
Parallel arrangement
In this system two separate parallel CO2 circuits operating between the ambient temperature on the high side and the intermediate and freezing temperature levels on the other sides. In order to obtain reasonable efficiency, the CO2 circuit that operates between ambient and freezing temperatures should have two‐stage compression with an intercooler. The following figure is a simple schematic of the parallel system solution.
Medium Temperature Freezer
Figure 2‐6: Schematic diagram of CO2 parallel system solution with two‐stage compression on the low temperature level
Booster system solution
In this system solution the refrigerant is expanded in two different pressure/temperature levels, medium and low. As can be seen in Figure 2‐7, the low stage compressor (booster) rejects the discharge gas into the suction line of the high stage compressor mixing with the superheated return vapour from the medium temperature level.
12
Figure 2‐7: Schematic diagram of CO2 booster system solution (TR2)
2.3 Heat Recovery System Solutions A refrigeration system without control to recover heat will operate in floating condensing where the condensing temperature follows the ambient temperature to a minimum condensing level which is usually 10°C. In this case, the heating needs in the supermarket are covered by district heating (DH) or a separate heat pump system (SHP). Figure 2‐8 is a schematic of system with floating condensing (FC) that rejects heat directly to the ambient. In case of R404A system it is most common that the condenser rejects heat to a coolant loop instead of directly to the ambient due to the charge minimization requirements.
Figure 2‐8: Schematic of the heat rejection side of a system with floating condensing arrangement‐ No heat recovery arrangement
Heat recovery system solutions in supermarkets are used mainly to heat the space air. Different system solutions for heat recovery are presented in the following sub‐sections.
13
2.3.1 Fixed head pressure (FHP) Figure 2‐9 shows the layout of FHP heat recovery. In this heat recovery solution the system operates in floating condensing and when heating is required the condensing pressure is elevated in order to provide the proper temperature for the heating system. As can be seen in the figure, the coolant extracts heat from the condenser of the refrigeration system and rejects the heat to the HVAC system before entering the dry cooler.
Figure 2‐9: Schematic diagram of the fixed head pressure heat recovery system
2.3.2 Desuperheater (DSH) Figure 2‐10 is a simple schematic of a system running with heat recovery in the de‐superheater. In this system solution heat is rejected in a de‐superheater which is installed before the air cooled condenser. Depending on the temperature level that can be reached in the heating system supply fluid, the system can provide heat to HVAC system or floor heating which requires lower operating temperature. The condensing pressure in the system is controlled according to the heating needs; the regulating valve, after the condenser in the schematics, controls the discharge pressure in the system according to the required capacity from the de‐superheater.
Figure 2‐10: Schematic diagram of the heat recovery at the de‐superheater
This heat recovery solution is viable in the systems operating with refrigerants that have high discharge temperatures. CO2 and NH3 systems have rather high discharge temperature compared to
14
R404A (50°C and 67°C compared to 30°C respectively, assuming isentropic compression between ‐5°C evaporation and 25°C condensing).
2.3.3 Heat pump cascade (HPC) Another concept of heat recovery from refrigeration system is the use of heat pump to extract heat from the condenser coolant at low temperature and transfer it to the HVAC system at high temperature levels. This system enables the use of rejected heat at the same time allows the refrigeration system to operate at relatively low condensing pressure.
Figure 2‐11: Schematic diagram of heat pump cascade heat recovery system
2.3.4 Heat pump cascade for subcooling (HPSC) Similar arrangement to the cascade heat pump solution is to connect the heat pump after the condenser in the refrigeration system so the refrigeration system operates at low condensing pressure when the ambient temperature is low and heating is needed. The heat pump at the same time recover the heat from the refrigeration system and provide further sub‐cooling to the refrigeration system and improve its efficiency. Figure 2‐12 is a schematic of a refrigeration system with heat pump at sub‐cooling.
15
Figure 2‐12: Schematic diagram of heat pump at sub‐cooling heat recovery system
16
3 Field measurements Refrigeration systems in two supermarkets with CO2 trans‐critical systems have been analysed for heat recovery. The data from these supermarkets have been collected through the online interface IWMAC (Iwmac, 2009). The electric power consumption of the compressor, pressure and temperatures at key points are recorded every 5 minutes. The important performance indicators are thereby calculated. Microsoft Excel have been used for data analysis and NIST reference properties have been used to calculate the properties of refrigerants through the REFPROP 7.0 (Eric W. Lemmon, Mark O. McLinden, & Marcia L. Huber, 2002).
Analyses of data from the field measurements have been used to generate assumptions and develop the different computer simulation models. Some of the parameters have been either taken directly or derived using the figures in the literature.
3.1 Calculations of heat recovery The supermarket is installed with measurements of temperature and pressure on the refrigerant circuits. Temperatures are also measured on the coolant loop. The main measurement points are indicated in Figure 3‐1. Power consumption of the compressors in the four units is measured separately. The estimation of heat recovery capacity is based on the parameters on the refrigeration side.
The mass flow of the refrigerant is estimated from the compressor manufacturer data at the measured conditions in the supermarket. The heat transferred across the evaporator, the condenser and the intercooler is then calculated as the product of mass flow rate of refrigerant and the enthalpy difference across the heat exchangers.
Heat loss in the oil cooler is calculated as the difference between the measured electrical power of the compressor and the calculated shaft power using the refrigerant mass flow rate and the enthalpy different across the compressor. Heat loss from the compressor body to the ambient is assumed as 7% of the measured electrical power input.
Total heat rejected from the refrigeration system is the sum of the condensers’, the inter‐coolers’ and the oil coolers’ capacities when connected to heat rejection loop.
3.2 CO2 transcritical system 1 The measurement is carried out on a supermarket located in the far north of Sweden. The refrigeration system is a parallel CO2 trans‐critical and the heat recovery arrangement is by using a heat pump in cascade to the refrigeration system. The system is a combination of the refrigeration system in Figure 2‐6 and the heat recovery system in Figure 2‐11. The system solution is referred to as trans‐critical system 1 (TR1).
3.2.1 System design and operation Figure 3‐1 is a schematic diagram of the system TR1. The refrigeration system is a parallel system with two circuits for low temperature level and two circuits for medium temperature level. The Low temperature units designated as FA1 and FA2 have two two‐stage compressors each. Medium temperature units KA1 and KA2 have four single‐stage compressors each. The compressors are from Dorin; TCDH 372B‐D for the low temperature and TCS373‐D for the medium temperature.
17
Oil cooler
Inter-CoolerOil cooler
+ -
Heat pump COP= 3.6
Q_max =300kW T_hp in= 13o C
- Pressure
-TemperatureKA1/KA2
FA1/FA2
T_con,b,o =18oC
45oC
-10oC-35oC
dTapp=5K
dTapp=5KdTapp=5K
dTsc=2K
dTsc=2K
dTsc=2K
Qoilcooler=15%
Qoilcooler=15%
dTsh=10KdTsh,ex=10K
dTsh,ex=15K
dTsh=10K
Figure 3‐1: Schematic diagram of the trans‐critical system 1 (TR1)
All four circuits reject heat to the common coolant loop which is cooled by dry cooler. The compressor oil cooler also reject heat to the same coolant loop. A heat exchanger is connected to the coolant loop before the dry cooler to recover heat. This heat exchanger is connected to 300kW heat pump, which is meant to deliver heat to the HVAC system of the supermarket. At the time of investigation the heat pump is not in operation, so all the heat is reject through the dry cooler.
It can be observed in the schematic that the coolant system at the condensers side rejects heat to the heat pump system via an additional heat exchanger and not directly to the heat pump evaporator. The decision of connecting the refrigeration system to the heat pump was taken at later stage of the project and it was technically easier to establish the connection via an additional heat exchanger. It might be also related to the separate legal responsibilities of the refrigeration and HVAC systems; so the systems are treated as two separate envelopes connected by a heat exchanger.
3.2.2 Existing mode of operation and selection of operating conditions for heat recovery with heat pump
At the time when the system has been analyzed the heat recovery system was not in operation due to technical problems which were not related to the solution concept. Therefore, the performance in the system has been analyzed in the refrigeration mode and then the system performance in heat recovery mode was possible to speculate and evaluate using the system performance and design data.
18
Figure 3‐2 shows the plot of total cooling capacity, average condensing temperature (T_condensing), common supply temperature of coolant brine (T_com,b,o) and total heat rejected in the system. The top line is the curve of total heat rejected by the system. This includes heat rejected from oil cooler and the inter‐cooler of the two‐stage units at the low temperature level. Oil cooler capacity was found to be rather constant over the year at about 8kW and intercooler capacity was found to be about 13kW.
Figure 3‐2: Monthly average of cooling capacity, compressor power, and heat rejected based on the measurements of trans‐critical system 1(TR1)
The common brine temperature of the heat rejection loop for the months of January and February in 2008 was not recorded. The brine temperatures are monthly averages. In the existing mode of operation the brine temperature is above 18oC only in July and August. The design inlet temperature of brine to the heat pump is 13°C. This requires that the minimum supply temperature of brine from the refrigeration system to be 18oC; this is assuming 5K temperature difference across the heat exchanger. In order to operate the heat pump during the heating season, the refrigeration system would have to operate at discharge pressure corresponding to that of August (58 bar). This point of selection is highlight in Figure 3‐2.
The selection criteria of operating conditions for individual units are also pointed out in Figure 3‐3 and Figure 3‐4.
0
50
100
150
200
250
Jan_
08
Feb_0
8
March_
08
April_0
8
May_0
8
June
_08
July_
08
Aug_0
8
Sept_0
8
Oct_08
Nov_0
8
Dec_0
8
Jan_
09
Feb_
09
March_
09
Month
Q,E
(KW
)
-10
-5
0
5
10
15
20
25
T (C
)
Cooling capacity Compressor power T_ambient T_com,b,o T_condensing
19
Figure 3‐3: Existing mode of operation of medium temperature unit (KA1) and selection of operating condition for heat recovery in TR1
From Figure 3‐3, the operating conditions for medium temperature unit to run the system in heat recovery mode is selected as pointed out by the ellipse in the figure. Corresponding to common brine outlet temperature of 18oC, condensing pressure is selected at 58 bars and the corresponding cooling COP of 3.2 in the heat recovery mode.
Figure 3‐4: Existing operating conditions of low temperature unit (FA1) and selection of operating conditions for heat recovery in TR1
0
10
20
30
40
50
60
70
80
90
100
Jan_
08
Feb_0
8
March_
08
April_0
8
May_0
8
June
_08
July_
08
Aug_0
8
Sept_0
8
Oct_08
Nov_0
8
Dec_0
8
Jan_
09
Feb_0
9
March_
09
Month
Q(k
W),T
(C)
0.0
0.5
1.0
1.5
2.0
2.5
3.0
3.5
4.0
4.5
5.0
CO
P
Cooling capacity Total heat rejected T_cond T_com,b,o COP
0
5
10
15
20
25
30
35
40
Jan_
08
Feb_0
8
March_
08
April_0
8
May_0
8
June
_08
July_
08
Aug_0
8
Sept_0
8
Oct_08
Nov_0
8
Dec_0
8
Jan_
09
Feb_0
9
March_
09
Month
Q(k
W),T
(C)
0.0
0.2
0.4
0.6
0.8
1.0
1.2
1.4
1.6
1.8
2.0
CO
P
Cooling capacity Total heat rejected T_com,b,o T_cond COP
20
From the above figure, operating conditions for heat recovery mode is selected corresponding to common brine outlet temperature of 18oC. Therefore low temperature units need to be operated at condensing temperature of at least 20oC at COP of 1.7. These operating conditions have been assumed to calculate the heat that can be recovered if operated in heat recovery mode.
3.2.3 Recoverable heat with heat pump solution From the above selections, the refrigeration system is set to operate at condensing temperature of 20oC on the heat recovery mode. The COP of medium and low temperature units are set to 3.2 and 1.7 respectively during the heat recovery mode. Using these values, compressor power, oil cooler capacity, and the total heat rejected is calculated. Total heat rejected is used as the heat source for the heat pump. Using the design COP of the heat pump, which is obtained from the heat pump manufacturing data (CIAT, 2010), total heat that is provided by the heat pump is calculated. This potential is compared with the maximum capacity of the heat pump, thus heat potential is limited to maximum capacity of the heat pump. Power consumption of the heat pump is estimated by dividing the total heat supplied by the heat pump by its COP.
Total heat recovery as calculated is presented in Figure 3‐5. To compare the heating potential and the maximum capacity of the heat pump, heat recovery was considered even for the warm months of June, July and August. Presented in the plot from bottom to top of the curve are power consumption of the compressors of the refrigeration system while operating without any heat recovery (E_ref_ref only mode), the power consumption of the compressors of refrigeration system when operating on heat recovery mode (E_ref_HR mode), total power consumption in heat recovery mode (E_tot_HR mode), Cooling capacity of the refrigeration system, COP of heating and heating capacity of the heat pump. The total power on heat pump mode (E_tot_HR mode) is sum of the power consumed by the compressors of the heat pump and the refrigeration system.
Figure 3‐5: Estimated heat recovery potential with heat pump system.
0
50
100
150
200
250
300
Jan_
08
Feb_0
8
March_
08
April_0
8
May_0
8
June
_08
July_
08
Aug_0
8
Sept_0
8
Oct_08
Nov_0
8
Dec_0
8
Jan_
09
Feb_0
9
March_
09
Month
Q,E
(kW
)
0
1
2
3
4
5
CO
P
E_ref_ref only mode E_ref_HR mode E_tot_HR modecooling capacity heat pump capacity COP heating
Maximum capacity of heat pump= 301kW
21
The flat curve for heat pump capacity from April to October is due to reaching the maximum capacity of the heat pump. The upper limit of COP of heating is limited by COP of the heat pump.
As it has been pointed out earlier, the results are an estimate combining the measurements with system requirements and manufacturer data. Since the heat pump was not in operation it was not possible to obtain the heating demand in the supermarket; therefore, the potential heat recovery was used instead of the real heating demand. However, the heating COP can be used from the above figures as a good indication of the system performance.
4.2 CO2 transcritical system 2 This supermarket is located near the city of Goteborg, which is at the western coast of Sweden.
3.2.1 System design and operation
Oil cooler
Ground heat source
+ +
Floor heating
-Pressure
-Temperature
Desuperheater Desuperheater
KAFA1/KAFA2
KA3
Figure 3‐6: Schematic diagram of trans‐critical system 2 (TR2)
This is a CO2 trans‐critical system with two booster units for the low and medium temperature levels (KAFA1&KAFA2) and a separate medium temperature circuit (KA3). Compressors in the medium temperature circuit, and high stage of booster units, are of TCS373‐D model from Dorin with swept volume of 12.6m3/h at 2900rpm. The booster compressors are of SCS362‐D model from Dorin with swept volume of 10.7m3/h at 2900rpm.
On the heat rejection side, all three units are connected with a de‐superheater before the air cooled condenser/gas cooler. The de‐superheaters are connected to a common brine loop which is used to transfer the recovered heat. The recovered heat is used for floor heating and HVAC system of the building. The compressors’ oil in this system is cooled by a separate air cooler.
22
The heat recovery system operates so as to maintain a certain supply temperature of the brine to the heating system, in the field measurement the average heating system supply temperature was about 35oC. The heat recovery capacity of individual unit is maintained by opening or closing the electronic valve connected to the refrigerant line after the gas cooler. This raises or decreases the discharge pressure thereby controlling the de‐superheater capacity. The supply temperature of the brine is also controlled by controlling the flow rate of brine to the de‐superheater. Supply of brine to each de‐superheater is controlled with the variable speed pump. The refrigerant line is also externally sub‐cooled from the borehole before the supply to the cabinets.
To calculate the heat recovery capacity on the refrigerant side the measurements of temperature and pressure before and after the de‐superheater is necessary. While measurements of pressure and temperature for key points on the refrigerant line were available from September 2008, the measurement of temperature at the exit of de‐superheater was available only from March 2009. The temperature measurement of brine in and out of the de‐superheater was available from October 2008. The average difference between the temperature of hot gas out of de‐superheater and that of the temperature of brine at the inlet of the de‐superheater for March is used to back‐calculate the hot gas temperature for the past months. The average temperature difference between hot gas and the brine inlet for the heat exchanger of the medium temperature unit KA3 was 4K, but for low temperature unit KAFA1 was 1K. Since the approach temperature of 1 and 0 is considered too low, which could be attributed to measurement inaccuracy, both values have been discarded and approach temperature of 4K was used for all the de‐superheaters. Therefore the temperature difference of 4K is added to the common brine inlet temperature to get the temperature of hot gas exiting the de‐superheater for the missing measurements. The procedure can also be followed in Figure 3‐7.
Figure 3‐7: schematic diagram of de‐superheater showing the assumed parameter in the heat recovery calculation in TR2.
23
3.2.2 Total heat recovery Table 3‐1 presents monthly average values of power consumption, cooling capacity, COP of refrigeration, total heat rejected, heat recovered and sub‐cooling with borehole. The heat recovered is the heat rejected in the de‐superheater. The percentage of heat recovery is calculated as the ratio of heat recovered in de‐superheater to that of the total heat rejected by the refrigeration system. The percentage of heat recovered through de‐superheating varies from 24 % to 35 %.
The sub‐cooling with the borehole is expressed as the percentage of cooling capacity to indicate improvement due to the sub‐cooling with borehole. The total heat rejected in the system is equal to the sum of cooling capacity and compressor power minus oil cooler capacity, borehole capacity and heat loss from the compressor body. The heat loss from compressor body is assumed as 7% of the total electric compressor power.
The heat recovery capacity from the refrigerant side has been compared to measurements from the HVAC side and the difference has been found to be about 4%, which is marginal for a system of such size.
Table 3‐1: Monthly average values of heat recovery with the supply temperature of brine to the heating systems of 35oC in TR2
24
The relationship between cooling capacity, heat recovery and condensing temperature can be observed in Figure 3‐8. Since heat recovery is only in the de‐superheater, in the plot heat recovery capacity is the de‐superheater’s (Q_desuperheater).
Description Oct‐08
Nov‐08
Dec‐08 Jan‐09 Feb‐09 Mar‐09
Apr‐09
Average outdoor temperature (o C)
10 6 4 3 2 5 11
Compressor power (kW) 57 59 57 54 55 56 55
Cooling capacity(kW) 183 177 174 163 165 168 176
Total COP 3.34 3.10 3.04 3.00 3.00 3.02 3.27
Total heat rejected (kW) 211 201 179 166 164 170 187
Q_desuperheater (heat recovered) (kW)
50 55 56 55 57 56 46
Q_borehole (sub cooler)(kW)
9 12 29 30 33 32 24
Q_oilcooler (kW) 11 11 12 11 12 11 10
% of Heat recovery = (Q_desuperheater/total heat rejected)
24% 27% 31% 33% 35% 33% 25%
Q_borehole as percentage of cooling capacity
5% 7% 17% 19% 20% 19% 13%
25
Figure 3‐8: Total cooling capacity, heat recovered and borehole capacity in TR2
The oil cooler capacity doesn’t change much over time. The evaporation temperatures are found to be rather constant over time; for low temperature unit at ‐35oC and medium temperature at ‐10oC. It is seen that the heat recovery capacity is increased during colder months of December to March compared to warmer months of October, November and April. Therefore condensing temperature during the colder months is slightly higher by 3oC and COP of the system drops slightly; 3 compared to 3.3 during the warmer months.
The capacity is increased by raising the condensing pressure, which is achieved by closing the electronic valve in the refrigerant line after the condenser. The temperature of the brine is maintained by regulating the flow rate of brine into the de‐superheater. It is more convenient to understand these controls on individual units over shorter period and has been discussed in the following sub‐sections.
3.2.3 Performance analysis of medium temperature unit (KA3) Figure 3‐9 presents the monthly average of cooling capacity, COP, condensing temperature, heat recovered and the borehole capacity.
0
10
20
30
40
50
60
Oct-08 Nov-08 Dec-08 Jan-09 Feb-09 Mar-09 Apr-09
Month
Q, (
kW),
T(C
)
1.0
1.5
2.0
2.5
3.0
3.5
CO
P
T_ambient Q_desuperheater Q_borehole T_cond COP
26
Figure 3‐9: Monthly performance in averages of medium temperature unit (KA3) in TR2
The heat recovery capacity increases, while the cooling capacity decreases. The increase in heat recovery capacity and increase in condensing temperature are directly proportional. This reduces the COP of the refrigeration system. The borehole sub‐cooling capacity is low during the period from October to January, while condensing temperature has been increasing during these months. Sub‐cooling from the ground source reduces the compressor power consumed and therefore increases COP. The increase in COP due to sub‐cooling can be observed in the months of February and March.
Heat recovery and control of medium temperature unit for 24 hours is presented in Figure 3‐10. The top curve is the percentage opening of the electronic valve in the refrigerant line after the condenser. The middle curve is the discharge pressure. The bottom curve is the heat recovered in the de‐superheater. It is seen that heat recovery starts at about 7 AM and stops at around 10 PM. The discharge pressure is increased after 7 AM to meet the heating demand. The pressure is increased by partially closing the electronic valve. The valve is almost 100% during the night time when there is no heating demand.
0
10
20
30
40
50
60
70
80
Oct-08 Nov-08 Dec-08 Jan-09 Feb-09 Mar-09 Apr-09
Month
Q,E
(kW
),T(C
)
2.0
2.5
3.0
3.5
4.0
4.5
5.0
CO
P
T_cond Cooling capacity Q_desuperheater T_ambient Q_borehole(subcooling) COP
27
Figure 3‐10: Operation of medium temperature unit over a period of 24 hours in TR2
4.2.4 Performance of booster unit (KAFA1) The monthly average of cooling capacity, COP and heat recovery capacity are presented in the following plot.
Figure 3‐11: Average monthly performance of (booster system) (KAFA1) in TR2
0
10
20
30
40
50
60
70
80
90
100
00:00 02:25 04:30 06:55 08:40 10:20 12:10 13:55 16:00 17:50 20:35 23:00
Tiime (march 24,2009)
Q(k
W),
P(ba
r), C
apac
ity(%
)
Electronic valve opening 2 per. Mov. Avg. (Discharge Presure) 2 per. Mov. Avg. (Q_desuperheater)
0
10
20
30
40
50
60
Oct-08 Nov-08 Dec-08 Jan-09 Feb-09 Mar-09 Apr-09
Month
Q,E
(kW
),T(C
)
1.0
1.5
2.0
2.5
3.0
3.5
CO
P
T_cond Q_desuperheater Q_borehole(subcooling) cooling capcacity COP
28
As can be seen in the plot, it is necessary to maintain condensing temperature of about 23oC in the colder months to recover the same or less amount of heat to that of October and April. This may be due to the drop in cooling capacity in colder months. The decrease in cooling capacity is due to the decrease in cooling demand in the medium temperature unit. The cooling capacity of low temperature unit is found to be rather constant.
The control of the booster system over a day period and heat recovery capacity is presented in Figure 3‐12. It can be seen that the discharge pressure is controlled within smaller range of variation than in the case of KA3. This could be accounted for the fact that the heat recovery capacity in the booster system is higher than in case of the medium temperature level system.
Figure 3‐12: Operation of booster unit (KAFA1) over a period of 24 hours.
3.3 Conclusions With the drive to reduce net energy consumption, heat recovery is gaining popularity in supermarket refrigeration systems. In Sweden most of the new supermarkets are installed with heat recovery systems. Field measurements of TR1 and TR2 did give insight on the performance of the systems in heat recovery mode. The work in this project also demonstrated the control of the system, especially TR2, and provided better understanding of the system behaviour.
In case of TR1, actual field measurements on the heat recovery system would have provided more information on the performance of the system, unfortunately, for technical reasons the heat pump system was not in operation.
For TR1, if the refrigeration system was operated without heat recovery, condensing temperature was as low as 13°C, where as to operate in heat recovery mode with heat pump the minimum
0
10
20
30
40
50
60
70
80
90
100
00:00 01:40 03:20 05:05 06:45 08:25 10:05 11:45 13:25 15:05 16:45 18:25 20:05 21:45 23:25
Time (March 24, 2009)
Q(k
W),P
(bar
), C
apac
ity (%
)
Electronic valve opening 2 per. Mov. Avg. (Discharge pressure) 2 per. Mov. Avg. (Q_desuperheater)
29
condensing temperature has to be 20°C. This puts burden on the refrigeration system by reducing the COP of refrigeration system. But on the other hand, large capacity of heat is available. The power consumption as calculated is for the maximum capacity; therefore, the high values of power consumption by the heat pump may be misleading. In reality the heat pump should be running in part load thus lower power consumption than presented.
From the field measurement of TR2, it was found that up to 30% of the total heat rejected can be recovered at a temperature of 35°C. The rest of heat is rejected to the ambient through the air cooled condenser. It is however seen that the amount of heat recovered is less compared to TR1 where the condensing temperature has to be raised above 20°C, which increases the pressure ratio thus the COP of refrigeration drops. The TR2 has higher efficiency mainly because there is the external sub‐cooling from the ground source.
One of the problems faced in this project was to get the reading at the point of interest in both the refrigeration side and on the heat recovery side. The comprehensive field measurement that provides information on the available heat on refrigeration system side and on the heating demand side is necessary to give better indication of system performance.
In reality the systems solutions vary in some form or other. This makes it difficult to compare different systems based on a certain performance criteria such as COP refrigeration system, COP of heating, power consumption heating and cooling capacity etc. Therefore computer simulation modelling is important to develop fair comparison of different systems.
30
4 Computer simulation modelling Modelling is an important part of this study because of the difficulty to compare systems in field measurements. Systems in real installations have different settings, operating conditions, capacities and requirements. It is more challenging when the comparison is for the simultaneous heating and cooling, which adds more variables to the systems under investigation. Also in the computer simulation models it will be possible to compare systems that do not exist in real installations using input variables from existing systems.
The assumptions for the different systems have been kept as practically similar as possible; therefore, the results of the comparisons should be treated as relative to the systems under comparison. Models are written using EES software, its basic function is to provide the numerical solution to a set of algebraic equations. It has many built‐in mathematical and thermo‐physical property functions for refrigerants (Klein, 2006).
Details of the assumptions used in the models can be found in Freléchox (Freléchox, 2009) and Nidup (Nidup, 2009). Some of the key assumption is that the cooling capacity of the refrigeration system at the medium temperature level is constant below 10°C ambient. Freezing capacity is assumed to be constant which agrees with data from the field measurements, this may be attributed to the use of covers on the freezing cabinets.
Below 10°C ambient, heating demand starts (Nidup, 2009) and it has been simulated for an average size supermarket in Sweden using CyberMart (Arias, 2005) by Jaime Arias. The generated heating demand values by CyberMart have been plotted against ambient temperature and the plot in Figure 4‐1: has been generated.
Figure 4‐1: Heating demand in average size supermarket in Sweden at different ambient temperatures
0
50
100
150
200
250
300
‐40 ‐30 ‐20 ‐10 0 10 20 30 40
Heating
dem
and (kW)
T,amb (C)
Heating demand in average size supermarket
31
As can be seen in the plot, heating demand starts when the ambient temperature is lower than 10°C and increases for lower temperatures. The start of the heating is an input variable to the calculations model but the capacities are dependent on the supermarket envelope, requirements and activities, etc.
4.1 System performance in floating condensingNo heat recovery Several systems have been analyzed but the main systems that will be presented in this chapter are the parallel trans‐critical CO2 system (TR1), conventional R404A and the booster CO2 system with low pressure receiver (TR3). Simple schematic of the R404A conventional is presented in Figure 2‐2. The booster system (TR3) is similar to the system presented in section 2.2.3 except for the additional receiver after the high stage regulation valve, schematic of the system TR3 is presented in Figure 4‐2.
Figure 4‐2: Schematic diagram of CO2 booster system solution with low pressure receiver (TR3)
In the TR3 system solution the pressure in the receiver is maintained at level reasonably higher than for the medium temperature/pressure level. Vapour is extracted from the receiver and flashed into heat exchanger which sub‐cools the liquid on one side and evaporates liquid droplets on the other. The main influence the receiver has is to improve the heat transfer in the evaporators due to the lower vapour content entering the evaporator and to increase the efficiency of the low temperature circuit. The influence of the low temperature circuit is highlighted in the process plot on the P‐h diagram in Figure 4‐3.
32
Figure 4‐3: P‐h diagram of the refrigeration process in the system TR3
TR1 system is similar to the solution presented in Figure 2‐6 in section 2.2.3 with an addition of mechanical sub‐cooling at the low temperature unit by the medium temperature unit, as can be seen in the following figure.
Figure 4‐4: Schematic of TR1 system with mechanical sub‐cooling
The systems have been simulated for the evaporation temperatures, ‐35°C and ‐10°C at the low and medium temperature levels respectively. Condensing is assumed to take place in air cooled condenser/gas cooler; this is a conventional solution for CO2 but not for R404A systems. The
33
minimum condensing temperature of the systems is assumed to be 10°C. Figure 4‐5 is a plot of the medium and low temperature COP’s of the three different systems in floating condensing mode.
Figure 4‐5: Medium and low temperature COP of different system solutions for different ambient temperatures.
As can be observed in the plot, the CO2 systems, TR1 and TR3, have higher COP than the R404A conventional at ambient temperatures lower than 20°C; this is mainly due to the presence of the brine loop in the R404A system and due to loss in COP for the CO2 systems when operating trans‐critically.
The R404A conventional system has higher low temperature COP than the CO2 systems, especially at high ambient/heat rejection temperatures. The TR3 system has similar freezer COP to TR1, this is mainly due to the influence of the receiver which is more prominent at high discharge pressures.
4.2 System performance in fixed head pressure (FHP) heat recovery In this solution for heat recovery the systems will have to operate with a coolant loop connected to the condenser and the HVAC system as sketched in Figure 2‐9. When heating is needed in the supermarket, ambient temperature is lower than 10°C, then the system will have to operate at high discharge pressure in order to provide the coolant with a temperature of 45°C to the heating system; accordingly, the system will have low refrigeration system COP’s in the heat recovery mode.
The COP of the TR1 and R404A conventional system in the FHP is plotted in the following figure. The system TR3 will have comparable COP’s to TR1, as can be observed in Figure 4‐5, therefore it will not be included in some of the following comparison figures.
0
1
2
3
4
5
6
-10 -5 0 5 10 15 20 25 30 35 40
CO
P (-
)
Ambient temperature (ºC)
Med & Low COP- Air cooled condenserTR1,med
TR3, med
R404A conv, med
R404A conv, Fr
TR1, Fr with SC
TR3, Fr
34
Figure 4‐6: Medium and low temperature COP of TR1 and R404A conventional system in the FHP heat recovery system solution.
As can be seen in the plot, the CO2 system will have much lower COP compared to the conventional R404A system, especially for the low temperature COP, due to operating at high discharge pressure. This indicates that the heating COP of CO2 in such system solution is quite low compared to R404A system especially that the CO2 systems have steeper COP lines than the conventional system and therefore more sensitive to increase in discharge pressure.
4.3 System performance in heat pump cascade (HPC) heat recovery When heat pump is connect in a cascade arrangement to the refrigeration system the system will have to provide the coolant to heat pump at lower temperature than in the FHP arrangement. Applying the conditions for the system TR1 that has been analyzed in the field measurements, Figure 3‐1, the refrigeration system is required to provide a coolant temperature to heat pump system of 13°C, consequently, it will have to operate at a condensing temperature of about 20°C.
The COP of the TR1 and R404A conventional systems in the HPC heat recovery mode is plotted in the following figure.
0
0,5
1
1,5
2
2,5
3
3,5
-10 -5 0 5 10 15 20 25 30 35 40
CO
P (-
)
T,amb [°C]
TR1 Med R404A conv. Med TR1 Fr R404A conv. Fr
35
Figure 4‐7: Medium and low temperature COP of TR1 and R404A conventional system in the HPC heat recovery system solution.
As can be observed in the plot, TR1 system has higher efficiency at the medium temperature level than the R404A when the system is operating in the heat recovery mode. However, the freezer COP for the TR1 system is lower for all the temperature range. Since the COP curves in the plot intersect, a clearer indication of the system performance is to calculate the annual energy consumption of each system solution which will presented in later section of this chapter.
The COP of the heat pump that is connected in cascade to the refrigeration system is obtained from the manufacturer data (CIAT, 2010) and plotted in Figure 4‐8.
0
0,5
1
1,5
2
2,5
3
3,5
4
-10 -5 0 5 10 15 20 25 30 35 40
CO
P (-
)
T, amb [°C]
Heat pump cascade-45C TR1 Med
R404A conv. Med
TR1 Fr
R404A conv. Fr
36
Figure 4‐8: COP1 of the heat pump in the TR1 system (CIAT, 2010)
For coolant outlet temperature from the heat pump evaporator of 7°C and to provide the heating system with 45°C supply the heat pump will have a COP of about 3,7.
The cooling capacities profile at the medium temperature level is 100kW below 10°C and reaches 200kW at 35°C. The cooling capacity at the low temperature level is constant and equal 35kW. Using the COP’s of the refrigeration system and the heat pump the annual energy consumption with the ambient temperature of Stockholm results in comparable energy consumption for both systems, slightly less for the R404A system.
4.4 System performance in heat pump cascade for subcooling (HPSC) heat recovery
Figure 4‐9 is a schematic of a booster system with heat pump connected after the condenser/gas cooler. In this solution the refrigeration system operates in floating condensing and for ambient temperature lower than 5°C the system operates at the minimum condensing pressure. The heat pump extracts the necessary heat from the refrigeration system and provides sub‐cooling to 7°C. The de‐superheater capacity is also recovered without controlling the condensing pressure, return temperature from the heating system is assumed to be 30°C.
0
1
2
3
4
5
6
‐10 ‐5 0 5 10 15
COP1
(‐)
Evaporator water outlet temp (C)
COP1 of a heat pump in HPC system
Cond water outlet temp=35C
Cond water outlet temp=45C
37
Figure 4‐9: Schematic diagram of booster system with heat pump at sub‐cooling for heat recovery
Compared to the HPC refrigeration/heat recovery arrangement, described in section 4.3, in the HPSC solution the refrigeration system operates at lower condensing pressure due to the direct heat rejection to the ambient and the heat pump unit will operate at lower evaporation temperature.
The energetic performance of this system solution in relation the other systems in this study will be presented in the following sections.
4.5 System performance in desuperheater (DSH) heat recovery Figure 4‐10 is a simple schematic of a booster system running with heat recovery in the de‐superheater. In this system solution heat is rejected only in the de‐superheater which is installed before the air cooled condenser. The discharge pressure of the system is controlled according to the required heat in the supermarket.
38
20 or 30C
Figure 4‐10: Schematic diagram of a booster system with heat recovery in the de‐superheater
Figure 4‐11 demonstrates how the system is being controlled in the de‐superheater heat recovery mode. As can be seen in the figure, when the discharge pressure increases the available heat for recovery in the de‐superheater increases. In this case the return temperature from the heating system is assumed to be 20°C and 5K approach temperature difference is assumed between the CO2 in the de‐superheater and the heating system working fluid.
Figure 4‐11: P‐h diagram of the heat recovery process in the DSH heat recovery
Figure 4‐12 is a plot of the system COP at diffenent ambient temperatures when the system is controlled for heat recovery in the de‐superheater. As can be seen in the plot, when the ambient temperature drops, right side of the pot, the COP increases due to lower condensing pressure. Heating demand starts at 10°C ambient and discharge pressure is raised to recover the required heat, consequently, the COP of the refrigeration system drops.
39
Figure 4‐12: Medium temperature level COP and heating demand for different ambient temperatures. Heating system return temperature is 20°C.
The range of flat line of COP is due to reaching the condensing temperature and the possibility to reject all the refrigeration system heat in the de‐superheater. The assumed 20°C return temperature from the heating system is rather low, a more realistic value that has would be 30°C and both cases have been investigated in the further calculations.
The following plot shows the energy consumption of the refrigeration system at different ambient temperatures for heating system return temperatures of 20 and 30°C. A flat line in the energy consumption curve is also observed for the 20°C return temperature case, however, in the case of the 30°C return the line is not flat due to the shape of the isotherm over the critical point, the isotherm for 35°C can be can be observed in Figure 4‐11.
Figure 4‐13: Energy consumption of the refrigeration system at different ambient temperatures. For heating system return temperature of 20°C and 30°C.
40
In order to evaluate the system performance in heat recovery mode the heating COP of the system is defined as the ratio between the heating demand, plotted in Figure 4‐1Error! Reference source not found., to the energy consumed to provide the heat, which is the difference between the energy consumption of the refrigeration system in heat recovery mode and floating condensing mode. This can be observed in the difference between the processes in the following plot.
Figure 4‐14: P‐h diagram of two processes; heat recovery and floating condensing modes.
The heating COP of the booster system with heat recovery in the de‐superheater is plotted in the following figure for the cases of 20 and 30°C return temperature from the heating system. The straight line in the plot is the heating COP of a conventional heat pump system, about 3,2 according to research work done at IUC‐SEK (Rogstam, 2010).
Figure 4‐15: Heating COP of booster system with heat recovery from de‐superheater. For heating system return temperature of 20°C and 30°C.
41
As can be observed in the plot the system with 30°C return temperature from the heating system has lower COP than a conventional heat pump system. The refrigeration system in these calculations is not controlled for sub‐cooling in the gas cooler when the system is running in the heat recovery mode. When the ambient temperature is low the heating demand is high and the discharge pressure is raised to recover heat, in this case the gas cooler can be operated to further cool the refrigerant before passing the expansion valve.
Running the gas cooler to cool the refrigerant down in the heat recovery mode has two main effects on the system performance, it increased the system COP which is a positive influence but it also reduces the available heat to recovery in the de‐superheater at certain discharge temperature; this is due to the smaller mass flow rate running in the system with sub‐cooling. The sub‐cooling influence is plotted on the P‐h diagram in the following plot.
Figure 4‐16: P‐h diagram of the sub‐cooling influence on system performance
The influence on system performance per degree of sub‐cooling is plotted in the following figure.
Figure 4‐17: Influence of sub‐cooling on system COP and de‐superheater capacity
42
Accordingly, in order to recover certain heat from the system with sub‐cooling, it has to operate at higher pressure than without sub‐cooling. For the case of recovering 55kW from the de‐superheater, the system with sub‐cooling (5°C condenser/gas cooler exit temperature) operates at higher discharge pressure than without sub‐cooling but still with higher COP due to the positive influence of sub‐cooling.
Figure 4‐18: P‐h diagram of the sub‐cooling influence on system performance, operating the systems to provide 55kW heat from the de‐superheater
The positive influence on system performance is more prominent at certain higher heating demand/higher discharge pressure range. This is due to the operation being near the flat or semi‐flat isotherm where the system can reject most of its heat in the de‐superheater. For the case presented in the following plot the COP of the system with sub‐cooling is 2,8 compared to 2,1 if the case when gas cooler is not in operation.
43
Figure 4‐19: P‐h diagram of the sub‐cooling influence on system performance, operating the systems to provide 115kW heat from the de‐superheater
The influence of heat demand on the system COP can be seen in the following plot, in this plot the system is without sub‐cooling in the gas cooler in the heat recovery mode. The system is controlled to provide the heating demand.
Figure 4‐20: COP of the medium temperature unit in the booster system as a function of heat demands. The case is for the system without sub‐cooling in the gas cooler.
The further sub‐cooling in the gas cooler will have the positive influence on the system COP up to a point where the heating demand will be high and the system will have to operate at elevated discharge pressure levels because part of the capacity is rejected in the gas cooler. Therefore, the gas cooler has to be by‐passed in order to have all the heat recovered in the de‐superheater. Figure 4‐21 is a plot of the system’s medium temperature COP for the cases with and without sub‐cooling in the gas cooler.
44
Figure 4‐21: COP of the medium temperature unit in the booster system as a function of heat demands. The cases are for systems with and without sub‐cooling in the gas cooler.
Therefore, the system has to be controlled by running the gas cooler up to a certain heating demand level and then the gas cooler is by‐passed otherwise the system will suffer in a COP loss as observed in the above figure. A proper control of the system will result in the COP presented in the following figure.
45
Figure 4‐22: COP of the medium temperature unit in the booster system as a function of heat demands. The system is controlled for sub‐cooling in the gas cooler and then by‐passed at high heating demand.
The energy consumption of the refrigeration system when controlled to produce the COP’s in Figure 4‐22 is presented in the following figure.
Figure 4‐23: Energy consumption of the refrigeration system at different ambient temperatures. For heating system return temperature of 20°C and 30°C and for 30°C with sub‐cooling in the gas cooler.
46
The heating COP of the systems is then calculated and is presented in the following plot. One more system has been added to the comparison, it is the system with heat pump at sub‐cooling (HPSC).
Figure 4‐24: Heating COP of booster system with heat recovery from de‐superheater. For heating system return temperature of 20°C and 30°C.
It can be observed in the plot that the system with heat pump at sub‐cooling has higher heating COP than using a separate conventional heat pump system. However, using the sub‐cooling in the gas cooler has high heating COP for ambient temperatures higher than about ‐15°C (heating demand of about 150kW).
The curves trends in the plot suggest that the system with heat recovery at the de‐superheater will have relatively low energy consumption when the heating demand is rather moderate. Sub‐cooling in gas cooler is essential to improve the system performance in heat recovery mode, controlling the capacity/operation of the gas cooler could be a good way to control how much heat can be rejected in the system.
4.6 Annual energy consumption calculations After developing the system performance in floating condensing and heat recovery mode the annual energy consumption for an average size supermarket in Sweden in the climate of Stockholm area is calculated and presented in the following figure.
47
Figure 4‐25: Annual energy consumption of different system solution for refrigeration and heat recovery for a Swedish average size supermarket in Stockholm climate conditions.
Additional systems that need to be pointed out are the ones with separate systems for refrigeration and heating. The refrigeration system in this case is operating in floating condensing and the heating is provided either by a separate heat pump system or district heating. The district heating system has been assumed to be equivalent to a heat pump with a COP of 1,8 based on the ratio of the prices of heating and electricity (STEM, 2009).
The system that is denoted as TR1 alternative is presented in the following schematic. It is a parallel system that has mechanical sub‐cooling, de‐superheater heat recovery and heat pump at the sub‐cooling of the medium temperature unit. The condensing pressure is allowed to float with the ambient reaching the minimum value of 10°C.
Figure 4‐26: Schematic diagram of TR1 alternative system
48
As can be observed in Figure 4‐25, the systems with heat recovery from de‐superheat and with heat pump connected to the sub‐cooling has comparable energy consumption to the having a separate heat pump system to provide the necessary heat. The system TR1 Alternative has the lowest energy consumption among all the compared systems. It can be observed that almost all alternatives have lower energy consumption (energy cost) than providing the necessary heat via the district heating system.
4.7 Conclusions Modelling is an important part of this study because of the difficulty to compare systems in field measurements. Systems in real installations have different settings, operating conditions, capacities and requirements. It is more challenging when the comparison is for the simultaneous heating and cooling, which adds more variables to the systems under investigation. Also in the computer simulation models it is possible to compare systems that do not exist in real installations using input variables from existing systems.
Calculation models for several refrigeration and heat recovery solutions have been built. The COP’s of the different systems have been compared in floating condensing and heat recovery modes and the annual energy consumption for the systems have been calculated for the climate of Stockholm.
A technique used to recover heat in conventional systems is to elevate the condensing pressure to an acceptable heat rejection temperature to recover heat from the refrigeration system. When applying the same technique on CO2 trans‐critical system solutions the COP in heat recovery mode is much lower than conventional systems because of the loss of COP in case of CO2 when operating at high discharge pressures.
When connecting a heat pump to the refrigeration system in a cascade arrangement the efficiency of the CO2 medium temperature units are slightly higher than conventional system in the same arrangement; however, the CO2 low temperature units have lower COP than conventional. The annual energy consumption of both systems is comparable, slightly lower for the conventional system.
When connecting the heat pump at the sub‐cooling side of the CO2 refrigeration system and recovering heat from the de‐superheater, the resulting heating COP of the system is higher than using a separate heat pump system, especially at low heating demand.
Recovering heat from the de‐superheater of the refrigeration system results in high heating COP for moderate heating demands compared to conventional heat pump. Proper control of the system must be applied by running the gas cooler for further cooling of CO2 after the de‐superheater then the gas cooler must be by‐passed when the heating demand reaches high values.
The systems with the lowest energy consumption in providing the required cooling and heating in an average size supermarket are the systems with recovering heat from the de‐superheater, heat pump at the sub‐cooling side, and what has been defined as TR1 alternative. The systems have similar or lower energy consumption than a refrigeration system running in floating condensing with a separate heat pump to provide the required heating demand.
The system with recovering heat from the de‐superheater may have the lowest installation cost because of the absence of the separate heat pump system. The booster system has a cost advantage
49
over the parallel solution due to the need to have a single control system package/unit for the booster system while two are needed for the parallel system solution.
50
5 Experimental measurements of CO2 heat pump test rig
5.1 System Layout Division of applied thermodynamics at Energy department KTH has a prototype of CO2 heat pump, this is vapour compression cycle based water to water heat pump with nominal heating capacity of 30KW. The schematic diagram of the system is shown in the following plot,
Figure 5‐1: Schematics of the heat pump at KTH
The four main components are the compressor, gas cooler, expansion device and evaporator. All the heat exchangers (evaporator, gas cooler, internal heat exchanger etc) utilized in this prototype are of plate heat exchangers. As illustrated in the above diagram HEx 1, HEx 2 and HEx 3 are the plate type gas coolers and are placed in series configuration. Heat exchanger 1 and 3 are dedicated for tap water heating (high temperature application), while heat exchanger 2 is reserved for the space/floor heating application. With the manipulation of valves any of the heat exchanger can be bypassed, so the system can be operated in various modes. Internal heat exchanger is located between the gas cooler and the expansion valve. Liquid receiver is placed after the evaporator on the low pressure side. Test rig is equipped with semi‐hermetic compressor from Dorin having 5.4 m3/hr swept volume at 1450 rpm. Coriolis type flow meter is used for CO2 mass flow measurement while on the water side Brunata energy meters are adopted. Temperature measurement was done with Pt 1000 element from Danfoss.
5.2 Working principle After the compressor, CO2 passes through the oil separator where the lubricant oil is removed from CO2 and then it flows through the gas coolers, in the gas cooler portion heat is transferred from CO2 to the city water by virtue of the temperature difference. The CO2 then flows through the internal heat exchanger, mass flow meter and expansion valve in its way to the evaporator. Hot water from the hot water tank is used as heating load for the evaporator. After the evaporator CO2 passes through the receiver to the compressor and in this way completes the cycle.
51
5.3 Methodology Heat pump was operated to heat up the water (to a set temperature) for space/floor heating and tap water heating application, system was operated at fixed evaporation temperature and gas cooler side pressure was varied by changing the compressor speed (from 1050‐1800 RPM) and adjustment of expansion valve. Reading for temperatures, flow rate, pressures were taken when the system get stable. These parameters were then used in Engineering Equation Solver (EES) to evaluate the overall system performance (COP, Heating capacity, mass flow, etc).
Direct and Indirect mass flow measurement
The CO2 mass flow was checked with coriolis type mass flow meter, this type of flow meter has bent tubes for the fluid flow. As the fluid flows through the meter these tubes get some twist and this degree of twist in turns gives the mass flow reading (Wikipedia, 2010).Working principle of this meter suggests some limitations in its application, like when the fluid flow is rapidly varying or at low mass flow rate it becomes hard to have accurate reading from the meter. Sensitivity of such a meter varies with operating region, density change rate etc.
Density of the CO2 changes a lot depending upon the gas cooler side pressure and CO2 temperature at gas cooler outlet. Figure 5‐2 shows the variaton of CO2’s density with the variation of gas cooler outlet temperature at different gas heater pressures. The red retangular shows the region where the flow was measured ( the temperature of CO2 at gas cooler outlet, where the coriolis mass flow meter located). As shown in the figure, the density of CO2 has rapid changes in the tested region, which may cause problem in the readings (accuracy) of coriolis mass flow meter. Therefore, the measured CO2 mass flow was counter checked with other means in the current study as well, namely:
Figure 5‐2 Variation in density of CO2 with gas cooler pressure and gas cooler outlet temperature
• Heat exchanger energy balance method • CO2 mass flow rate from the compressor manufacturer’s data
52
HEx 1HEx 2HEx 3
IHEx
Evaporator
Expansion valve
Rec
ieve
r
Com
pres
sor
HWWater Heater
City waterCO2
HW
1246 5 3
ab
cd
ToutTin
7 8
Figure 5‐3 Schematics of heat pump with labels
Mass flow from meter reading
CO2 mass flow can directly be read from the C‐Type coriolis mass flow meter, mounted before the expansion valve.
Mass flow from energy balance
Energy balance on the gas cooler side is utilized to calculate the CO2 mass flow indirectly and the following expressions are adopted:
2 Heat lost by co Heat gained by water=
∆ ∆
∆
∆
Mass flow from compressor data
Compressor manufacturer provided us with the information of swept volume of the compressor and with the information of temperature and pressures of CO2 at the compressor suction, mass flow can be indirectly calculated as follows;
53
2
volumetric
s
in
volumetric efficiency of compressor
V swept volume flow rate of the compressordensity of co at compressor inlet
η
ρ
=
==
The volumetric efficiency of the employed compressor can be obtained by manufacture data fitting as shown in the following diagram.
Figure 3 compressor data fitting for volumetric efficiency
20.0081* 0.1593* 1.1243volumetric PR PRη = − +
where PR pressure ratio across the compressor=
Furthermore based on the different mass flows that calculated in above mentioned methods, different COPs can be calculated accordingly.
y = 0.0081PR2 ‐ 0.1593PR + 1.1243R² = 0.9776
0
0,2
0,4
0,6
0,8
1
0 1 2 3 4 5 6
η volum
etric
Pressure ratio
54
5.4 Testing results
5.4.1 Operating conditions Heat pump was operated to heat up the city water from 13 oC to 60 oC at an evaporation temperature of 0 oC, the CO2 temperature at the gas cooler outlet was maintained at about 33 oC.
5.4.2 Overall system performance Figure 4 shows the variation of CO2 mass flow rate with the variation of the gas cooler pressure, the results shows a good agreement between the direct (from corilois meter) and indirect (from energy balance and compressor data) reading.
Figure 5‐4 Variation of CO2 mass flow rate plus water and CO2 temperatures with gas cooler pressure
Figure 5‐5 shows the variation flow rates (for water and CO2) with the gas cooler side pressure. Water flow rate increases with the increase in the gas cooler side pressure.
Figure 5‐5 Variation of water and CO2 flow rates with gas cooler side pressure
Approach temperature (difference in CO2 and water temperatures at gas cooler inlet) is an important parameter, the high temperature glide in the gas cooler helps in achieving low approach value (Nekså, 2000). With the prototype at KTH the approach temperature was in the range from 15‐25 oC and its variation is shown in Figure 5‐6 with gas cooler side pressure.
0
0.02
0.04
0.06
0.08
0.1
0.12
0.14
0.16
7500 8000 8500 9000 9500
co2 mass flow rate
Gas cooler side pressure (KPa)
co2 mass flow rate v/s gas cooler side pressure
co2 mass flow direct reading
co2 from compressor data
co2 from energy balance
00,10,20,30,40,50,6
7500 8500 9500
Water / CO2 flo
w rate
Gas cooler side pressure (KPa)
Water and CO2 flow rates
water flow rate
co2 flow rate
Water m3/hrCO2 kg/sec
55
Figure 5‐6 Variation of approach temperature with gas cooler pressure
Figure 5‐7 shows how the heating capacity and the compressor power vary with the variation of gas cooler side pressure at constant evaporation pressure. As it is clear from the results heating capacity initially increases with the increase in gas cooler side pressure and eventually reaches to a maximum value, once this maximum point is crossed, the further increase in pressure no longer adds in the heating capacity.
Figure 5‐7 Variation of heating capacity and compression work
The effectiveness of the gas cooler is calculated from the ratio between the actual enthalpy drop of the CO2 and the maximum possible enthalpy drop. Figure 5‐8 shows the variation of the gas cooler effectiveness with the variation in gas cooler pressure. The following expression was used for the calculation,
1 6
1 6
Where h6water in
is the enthalpy of CO2 with water inlet temperature.
10
15
20
25
30
7500 8000 8500 9000 9500Approach Tem
perature (C)
Gas cooler side pressure (KPa)
Approach temperature v/s Gas cooler sider pressure
Approach temperature
Linear (Approach temperature)
0
5
10
15
20
25
30
7500 8000 8500 9000
capacity, w
ork (KW)
Gas cooler side pressure (KPa)
heating capacity and compressor work
Compressor work
heating capacity
56
Figure 5‐8 Gas cooler effectiveness v/s gas cooler side pressure
Figure 5‐9 shows the trend for the COP with the variation of gas cooler side pressure at contant evaporation temperautre (0°C). COP incresases initially with the gas cooler sider pressure and reaches a maximum value and then decreases again.
Figure 5‐9 Variation of COP with the gas cooler side pressure
5.4.3 Results for the compressor part Flow scheme and energy losses associated with compression process
CO2 enters the compressor from the suction side and first flows through the motor portion (cools down the motor, refrigerant absorbs heat), after flowing through the motor CO2 passes through the compressor and get compressed. After the compressor, oil separator removes the lubricant oil from CO2, this oil is cooled down by the oil cooler, which cools by water, and used again.
00.10.20.30.40.50.60.70.80.91
8000 8500 9000
Gas cooler effectiveness
Gas cooler pressure (KPa)
Gas cooler effectiveness
Gas cooler effectiveness
0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
7500 8000 8500 9000 9500
COP
Gas cooler side pressure (KPa)
COP v/s gas cooler pressure
COP direct reading
COP compressor data
57
Figure 5‐10 Flow Scheme in semihermetic compressor for transcritical operation (Fornasieri E, 2010).
During its flow through the motor CO2 gains some heat (by cooling the motor) and this energy is recovered by CO2. Water is used as a cooling media for oil cooler to cool down the compressor oil and then the cooled compressor oil is pumped back to compressor, which creates heat losses. This is called oil cooler losses in the current study. Furthermore, there is also a heat loss from compressor case to the surroundings as well and these two are the overall losses associated with the compression process.
The energy balance across the compressor can be expressed in the following equation accordingly.
Electricity consumed by compressor = Energy used in the real compression work + Oil cooler losses + Heat losses from the compressor body+ Energy gain by CO2 in motor
Isentropic efficiency of compressor
The isentropic efficiency of the compressor is calculated with following expression,
Where, hCO2out
is = Isentropic enthalpy of CO2 at the exit of compressor.
hCO2 in mid
= Enthalpy of CO2 at mid position (at compressor suction after motor portion).
hCO2out = Actual enthalpy of CO2 at the exit of compressor.
Figure 5‐11 shows the variation of volumetric and isentropic efficiencies against the gas cooler pressure. The results show that volumetric efficiency decreases with the increase in gas cooler pressure while isentropic efficiency shows almost no variation with gas cooler side pressure. High isentropic efficiency may be due to low pressure ratio and
58
Figure 5‐11 Volumetric and isentropic efficiencies with gas cooler pressure
Figure 5‐12 Compressor power consumption
Figure 5‐12 shows the variation in the compressor work with the gas cooler side pressures while Figure 5‐13 shows the temperatures and mass flow for the CO2 across the compressor. The pressure fo the CO2 on the gas cooler side was regulated by varying the compressor speed and adjustment in expansion valve, the highest value was obtained at maximum speed of 1800 RPM.
Figure 5‐13 CO2 mass flow rate and temperatures v/s gas cooler side pressure
Heat losses in the oil cooler
Energy lost by oil= Energy gained by cooling water
0%20%40%60%80%100%
65 75 85 95Volum
ertic efficiency
Gas cooler side pressure (bar)
volumetric efficiency
0%20%40%60%80%100%
65 85 105
Isen
trop
ic efficiency
Gas cooler side pressure (bar)
Isentropic efficiency
0
2
4
6
8
65 75 85 95
Power con
sumption (KW)
Gas cooler side pressure (bar)
compressor work
compressor work
59
Figure 5‐14 Oil cooler water flow rate and temperatures v/s gas cooler side pressure
Water flow rate in the oil cooler along with the tmeperatures are shown in Figure 5‐14, while the energy losses in the oil cooler are shown in Figure 5‐15. Energy losses are represneted in in terms of energy value (in KW) and as a percentage of total electric power consumption of the compressor.
Figure 5‐15 Oil cooler losses with gas cooler side pressures
Energy gain by CO2 in the motor portion
Temperature sensors measures temperature of CO2 at motor inlet and after the motor portion, while mass flow meter provides readings for the flow rate of CO2. Energy gained by the refrigerant in the motor portion was calculated by using the following expression,
Where, hCO2 in mid
= Enthalpy of CO2 after the motor portion.
hCO2 in = Enthalpy of CO2 at entry to the motor.
60
Figure 5‐16 Energy gained by CO2 in the motor portion v/s gas cooler side pressure
Heat losses from the compressor body
Radiation heat losses from the compressor body are investigated indirectly with the information of electric power consumption, oil cooler losses and energy gain by CO2 in motor portion by using the following relationship.
Where, Pel= Electrical power consumed by the compressor, and
is the real compression work.
Figure 5‐17 shows the trend for the heat losses from the compressor body (both in KW and as percentage of electric power consumption). This scatter is not having good trend and most of the points are even indicating negative heat losses from the compressor body which may due to the influence of the cold part of the compressor body, who is gaining heating from surrounding instead.
Figure 5‐17 Heat losses from compressor body with the gas cooler pressure
The computer simulation model and the heat exchanger performance evaluation
61
In parallel with the experimental work, a computer model of the test rig is also built in EES for
• Analyzing the CO2 heat pump’s performance
• Evaluating the heat exchanger performance
• Discussing system design issues
• Evaluating compressor performances
The operating window of the computer model is shown in figure 18.
Figure 18 operating window of EES computer model
As shown in the figure above, the testing results from the experimental work on the test rig are adopted as inputs to the computer model. The heat exchanger performance, heat balance and the thermodynamic performance of the heat pumps are then calculated. Furthermore the temperature profile of the water heating process in the gas cooler side will be plotted accordingly in a T‐h diagram for every testing condition as shown in figure 19.
62
Figure 19 T‐h diagram of one testing condition
The thermophysical properties of supercritical carbon dioxide heat pump and its heat exchangers UA value calculation.
The thermophysical properties of supercritical carbon dioxide will have sharp variations near its critical point, which is also the working region of heat pump’s heat recovering process for water heating. Therefore, the thermophysical properties of supercritical carbon dioxide needs to be carefully examined when one analyzing the performance of heat exchangers in the heat recovery process of a CO2 heat pump, due to its significant influence on the behavior of both the gas cooler and the internal heat exchanger (IHX) in the heat transfer process.
The specific heat (Cp), which is the main factor that influences the supercritical carbon dioxide’s temperature profile in heat pump’s gas coolers and IHX, is plotted as a function of the temperature for different pressures in the following figure (Fig. 20).
63
Figure 20 Specific heat of supercritical carbon dioxide vs. temperature at different pressures
It can be noticed from the figure that the specific heat of the supercritical carbon dioxide has more obvious changing dramatically with a peak value, when the pressure gets close to the critical pressure. Furthermore, it may also be noticed that the temperature corresponding to the peak specific heat value is increasing with increasing pressure.
As a comparison the Cp values of water from 15 °C to 90 °C (1 bar) are also plotted in the following figure.
Figure 21 Specific heat of water vs. temperature
0
5
10
15
20
25
30
35
40
45
50
10 20 30 40 50 60 70
Temperature (ºC)
CP(
kJ/k
g•K
)
P=7.4 Mpa
P=7.5 Mpa
P=8.0 Mpa
P=9.0 Mpa
P=10.0 Mpa
P=11.0 Mpa
P=12.0 Mpa
10 20 30 40 50 60 70 80 900
5
10
15
20
25
30
35
40
45
50
Twater
cpw
ater
(kJ/
kg K
)
1bar water Cp1bar water Cp
64
It can be noticed from the figure that the specific heat of the water is almost “constant”, compared to the specific heat changing of supercritical carbon dioxide near its critical point. Due to the variation in specific heats, the shape of the temperature profiles in the heat exchangers will be greatly influenced. It may also cause pinching1 in the heat exchanger. Therefore, this effect should be carefully examined when one evaluating the heat exchangers of carbon dioxide heat pumps. Furthermore, the traditional LMTD method and ε‐NTU method may not applicable for heat exchanger dimensioning.
In the computer model, the gas cooler are divided into small sections and the thermodynamic property of CO2 for every section is assumed to be constant. The LMTD method is therefore applied for every small section of heat exchangers to calculation the UA value for the section. The total UA value of the whole heat exchanger is then the integrated result of the results of all the small sections.
Figure 21 shows the change of calculated UA value against the number of divided sections of heat exchangers for one testing condition. It can be noticed from the figure that calculated UA value is changing with the increase of number of calculated heat exchanger section until the division of heat exchanger calculated section reaches a certain number (i.e. when the thermophysical values of supercritical CO2 can be assumed to be constant).
Figure 22 calculated heat exchanger UA values vs. number of divided heat exchanger sections
5.4.4 UA value with the flow rate The UA value of the gas cooler have also been tested by Wilson plot method, so fluid flow on one side (CO2 or water) was kept constant, while the other fluid is varied to see the effect of UA value. As we know the temperatures on both ends of the heat exchanger so we can calculate the LMTD, we know the heat capacity of the heat exchanger which helps in the calculation of the UA value.
The overall thermal conductance is calculated by following expression,
* ________________________________________________(10)mQ UA θ=
1 Pinching is the minimum temperature difference inside a heat exchanger , which limited the heat exchanger
kW/K
65
1 2
1
2
___________________________________________(11)ln
mT T
TT
θ Δ − Δ=
⎛ ⎞Δ⎜ ⎟Δ⎝ ⎠
Figure 5‐23 UA v/s water volume flow rate
Figure 5‐23 shows the variation in the UA value with variation in the water flow rate, the scatter is drawn for various compressor speeds and water flow was varied in each individual set while CO2 flow rate was held as constant.
Figure 5‐24 UA value v/s CO2 mass flow rate
Figure 5‐24 shows the variation of UA value with the variation of the CO2 mass flow rate, the scatter shows a good linear trend between the two plotted parameters.
5.5 Comparison with other studies Different researchers evaluated CO2 heat pump performance in various modes and with different arrangement of the components; however the main objectives were to optimize the system performance for tap water heating and space heating application. Most of the previous studies utilized counter flow type tube in tube type heat exchangers or shell and tube type heat exchangers for gas cooler and evaporator. Different researchers used different operating conditions, different system layouts and operating conditions makes the comparison task really difficult. Following case studies are use for the performance comparison of KTH heat pump.
0.7
0.75
0.8
0.85
0.9
0.95
0 0.2 0.4 0.6 0.8 1
UA (KW/K)
V . (m3/hr)
At 1450 RPM
At 1400 RPM
At 1350 RPM
At 1300 RPM
UA value v/s water flow rate at various speeds
0
0.2
0.4
0.6
0.8
1
1.2
0.05 0.07 0.09 0.11 0.13 0.15
UA value v/s co2 flow rate
UA value v/s flow rate
m_dot CO2 (kg/sec)
UA (KW/K)
66
5.5.1 Case study 1 Title: A carbon dioxide domestic hot water heat pump with double wall plate heat exchanger gas cooler (Fornasieri E, 2010). Objective of the study Heat pump system was operated to test the heat exchanger performance and overall system performance.
Component details
• Semihermetic piston type compressor with swept volume of 4.3 m3/hr. • Double wall plate heat exchanger used as gas cooler. • Plate type heat exchanger was used as evaporator. • Liquid receiver placed after the evaporator. • Electronic expansion valve controlled with stepper motor was used as expansion device.
Results
For heating up water from 14.7 oC to 60 oC at evaporation temperature of 1.8 oC and with gas cooler outlet temperature (for CO2) of 22.4 oC the COP was quoted as 3.8. The heating capacity of the system was 20.6 KW.
Figure 5‐25 shows the variation of thermal conductance (UA) with the variation of CO2 mass flow while water flow was kept constant.
Figure 5‐25 UA value v/s CO2 mass flow rate (Fornasieri E, 2010).
5.5.2 Case Study 2 Residential CO2 heat pump system for combined space heating and hot water heating (Stene, 2005).
Goal: performance investigation for space heating and tap water heating application.
System capacity 6.5 KW.
67
Figure 5‐26 Schematic diagram of the prototype
Components detail
• Compressor Hermetic two-stage rolling piston • Lubricant—polyalkylene glycol (PAG) • Evaporator Counter-flow single-pass tube-in-tube HX—stainless steel • Tripartite gas cooler Counter-flow single-pass tube-in-tube HX—stainless steel • Suction gas heat exchanger Counter-flow single-pass tube-in-tube HX—stainless steel
Results
The system was tested for only space heating, only district water heating and for the combined application of space heating and hot water heating cases. In the combined mode the evaporation temperature was ‐5 oC, tap water was heated from 6.5 to 60 oC and with various percentage of district heating capacity ratios. The results with various settings for hot water temperatures and with different space heating temperature ranges are summarized in the figure below. The measured overall isentropic efficiency for the prototype rolling piston compressor ranged from about 0.52 to 0.55 at 6000 rpm.
Figure 5‐27 Variation of COP with hot water temperature and DHW capacity ratio
68
5.5.3 Case Study 3 CO2 Heat pump water heater: Characteristics system design and experimental results (Nekså, 2000).
Goal: Performance evaluation of CO2 heat pump for tap water heating application.
Nominal heat capacity of 50KW and study was done during 1998.
Figure 5‐28 Detail of the prototype heat pump
Component details
• Old sabroe compressor from 1927 (one cylinder open type reciprocating). • Co-axial tube in tube counter flow heat exchanger was used as gas cooler. • Heat exchanger and tubing was made from stainless steel. • Evaporator was a plate in shell type heat exchanger • Expansion valve was regulated pneumatically by a computer according the required
conditions.
Results
Figure 5‐29 Variation of COP with the variation of evaporation temperature
Tap water was heated from 8 to 60 oC and at evaporation temperature of 0 oC and COP value of 4.3 was achieved. COP varies with the evaporation temperature and at evaporation temperature of ‐20 oC it was 3.
69
5.6 Conclusions In the current study, a CO2 heat pump with three gas coolers is built at the laboratory of Applied Thermodynamics and Refrigeration Division at KTH. All the heat exchangers of the test rig are plate heat exchangers. The test rig has a nominal heating capacity of 30 kW and it is built to
o Study the heat recovery performance of the heat pump gas coolers o Study the compressor performances o Provide experimental data and correlations to improve the computer simulation
model
A coriolis type mass flow meter is used to measure the CO2 mass flow, due to the dramatic change of CO2 thermophysical properties in the test region, the accuracy of direct mass flow metering is double checked with two indirect methods, namely heat exchanger heat balance and compressor data.
In general, the indirect method shows a good agreement with direct method in mass flow reading. The gas cooler effectiveness is increasing with increasing gas cooler pressure.
At certain evaporation temperature, both heating capacity and compressor work are increasing with increase gas cooler pressure, therefore, the heat pump COP shows an optimum at certain gas cooler pressure for a certain operating condition.
At 0°C evaporation temperature for instance, the test rig achieves an optimum heating COP of 4 (water is heated from 13°C to 60°C).
The compressor test shows:
Both volumetric efficiency and isentropic efficiency of the tested compressor (Dorin) are around 80%, while volumetric efficiency is decreasing with increase of gas cooler pressure and isentropic efficiency is approximately constant.
Oil cooler losses are increasing with increasing gas cooler pressure (compressor work), while the percentage of oil cooler losses in total compressor power consumption is decreasing.
In general the oil cooler losses are account for about 4% of the total compressor energy consumption. The heat gain of the refrigerant due to the compressor motor accounts for about 15% of the total compressor energy consumption.
Based on the experimental data, a computer simulation model is built to calculate and further analyze the test rig performance. The simulation results of the heat exchanger UA value shows that due to the thermophysical property change of supercritical carbon dioxide in contrast with relative constant value of water, the traditional LMTD or e‐NTU method are not suitable for calculating the heat exchanger UA value. Instead, the heat exchanger should be divided into several small sections to integrate its overall UA value.
The simulation results also show the importance of getting right dimension relation among the three gas coolers to ensure a good heat pump performance, this should be further investigated.
70
6 Overall conclusions Three work packages have been fulfilled in this project: field measurements analysis, computer simulation modelling and experimental evaluation. Two supermarket installations have been studied in the field measurements work package, several combinations of refrigeration and heat recovery arrangements have been tested in the computer simulation modelling, and a simplified system with key components and complete instrumentation have been built and tested in the laboratory.
The field measurements demonstrated the performance of the system and gave better understanding of the system’s behaviour and its control. It also provided key input parameters for the computer simulation modelling. The difficulty in estimating the heating demand in the real installations and the difficulty in comparing real installations stressed the need to use computer simulation modelling to evaluate and compare the different system solutions.
The experimental evaluation tested key components and control of the CO2 system in heat pump mode, overall system evaluation has been also performed. The methods used to estimate the mass flow rate in the field measurements have been verified, the compressor is key components where an extensive analysis have been performed. Good agreement between the actual mass flow measurements in the test rig and the estimation from the compressor manufacturer data has been observed.
In the computer simulation modelling , the systems with the lowest energy consumption in providing the required cooling and heating in an average size supermarket are the systems with recovering heat from the de‐superheater, heat pump at the sub‐cooling side, and what has been defined as TR1 alternative. The systems have similar or lower energy consumption than a refrigeration system running in floating condensing with a separate heat pump to provide the required heating demand.
The CO2 trans‐critical system with heat recovery from the de‐superheater showed good cooling COP and high heating COP at moderate heating demands in an average size supermarket in Sweden. It may have lower installation cost than the systems with comparable energy consumption due to the absence of the heat pump in the system and the need only for a single control package/unit in the system.
71
7 References
Arias, J. (2005). Energy Usage in Supermarkets Modelling and Field Measurements. Unpublished Doctoral thesis, Royal Institute of Technology, Stockholm.
CIAT. (2010). CIAT Water‐cooled chillers. Available: www.ciat.com [2010, June]. Eric W. Lemmon, Mark O. McLinden, & Marcia L. Huber. (2002). REFPROP: Reference Fluid
Thermodynamic and Transport Properties (Version 7). Fornasieri E. (2010). A corbon dioxide domestic hot water heat pump with double wall plate heat
exchanger gas cooler. Paper presented at the 9th IIR Gustav Lorentzen conference Freléchox, D. (2009). Field measurements and simulations of supermarkets with CO2 refrigeration
systems. Unpublished MSc, Royal Institute of Technology (KTH), Stockholm. Iwmac. (2009). Centralised operation and surveillance, by use of WEB technology
(http://www.iwmac.no/english/). Klein, S. A. (2006). Engineering Equations Solver (EES). Madison, USA: F‐Chart Software. Lundqvist, P. (2000). Recent Refrigeration Equipment Trends in Supermarkets: Energy Efficiency as
Leading Edge. Bulletin of the International Institute of Refrigeration(5). Madsen, K. B., Villumsen, J., & Nielsen, J. (2003, August 2003). Implementation of CO2/Propane
Walk‐In Storage Rooms in a McDonald's Restaurant. Paper presented at the 21st IIR International Congress of Refrigeration, Washington, D.C., USA.
Minea V. (2007). Energy Efficiency of a Supermarket Refrigeration/Heat Recovery System with Secondary Fluids. Paper presented at the International Congress of Refrigeration, Beijing, China.
Nekså, P. (2000). CO2 heat pump systems. International Journal of refrigeration. Nidup, J. (2009). Investigation of Heat Recovery in Different Refrigeration System Solutions in
Supermarkets. Unpublished MSc, Royal Institute of Technology (KTH), Stockholm. Reindl D. T, & Jekel T. B. (2007). Heat Recovery in Industrial Refrigeration. ASHRAE Journal, 49(08). Rogstam, J. (2010). Personal communication, www.iuc‐sek.se. Stockholm: IUC‐SEK. Sawalha, S. (2008). Carbon Dioxide in Supermarket Refrigeration. Unpublished Doctoral thesis, Royal
Institute of Technology, Stockholm. Sjöberg, A. (1997). Covering of a Cabinet in Supermarkets. Unpublished MSc, Royal Institute of
Technology, Stockholm, Sweden. STEM. (2009). Energy Indicators 2009, Follow‐up of Sweden’s energy‐policy objectives: Swedish
Energy Agency. Stene, J. (2005). Residential CO2 heat pump system for combined space heating and hot water
heating. International Journal of Refrigeration, 28(8), 1259‐1265. Wikipedia. (2010, June). Mass flow meter. Available: http://en.wikipedia.org/wiki/Mass_flow_meter. Wulfinghoff D. R. (1999). Energy Efficiency Manual. Energy Institute Press, Maryland, USA.
