ADVANCED COMPRESSOR DESIGN AND VARIOUS SYSTEMS FOR COMMERCIAL APPLICATIONS WITH CO 2

Transcription

1 ADVANCED COMPRESSOR DESIGN AND VARIOUS SYSTEMS FOR COMMERCIAL APPLICATIONS WITH CO 2 Oliver JAVERSCHEK (a), Dr. Günter DITTRICH (b), (a) Bitzer Kuehlmaschinenbau GmbH, Eschenbruennlestr.15, Sindelfingen, 71065, Germany, +49.(0) , javerschek@bitzer.de (b) Bitzer Kuehlmaschinenbau GmbH, Industriestr 48, Schkeuditz, 04435, Germany, +49.(0) , dittrich@bitzer.de Abstract This paper raises the topic how CO 2 is applied as refrigerant in refrigeration and how it could further be applied. It highlights how an advanced compressor design influences the application of CO 2 with respect to strength and safety, reliability and efficiency. Furthermore various system solutions with CO 2 are introduced and a discussion of their advantages and disadvantages, along with a comparison of system COP s is provided. In conclusion, the discussed issues will be analysed from the application engineering point of view. 1 Introduction CO 2 is an ongoing trend in refrigeration. Emissions of mostly used HFCs have a 1300 to 3300 times higher short term contribution to global warming than CO 2. Therefore, where refrigerant containment is somewhat questionable, CO 2 can be a solution for the future. Additionally the natural substance offers the possibility for differentiation from conventional design and a green image. In countries with a high taxation on HFCs, legislation is an additional driving force. In Denmark e.g. the tax value for 1 kg R404A of 588 DKK (approximately 80 ) is in discussion. The final passage of the new taxation by the Danish parliament is intended for 29 th of May Nonetheless, efficiency should not be neglected by introducing a refrigerant with a low GWP, especially because with a low GWP the efficiency becomes even more important for the assessment of refrigeration systems on basis of the TEWI method. Besides the efficiency, also strength, safety and reliability require an advanced compressor design. Focussing on the system design, increased efficiencies and reduced complexities are the aim. 2 Advanced compressor design Considering the thermodynamic properties of CO 2, it is obvious that unfavourable conditions with respect to the compressor design appear with higher discharge and gas cooler outlet conditions. These require high discharge pressures and cause low liquid fractions inside the two-phase area after the process of expansion. High discharge and standstill pressures on the suction side are a challenge concerning strength and safety aspects. Another challenge is the reliability of compressors while facing high mechanical loads for the drive gear and lubrication under unfavourable conditions. At the last but not at the least, trans-critical CO 2 applications require efficient compressors at full load and part load conditions together with a smart system and heat exchanger design to enable this technology to outperform HFC baseline systems in moderate climate zones. All this already highlights that compressors for trans-critical applications require a completely independent design and allow no real synergies to standard compressors. Nonetheless, a good compressor design considers as many proven assemblies from a related compressor range as possible. The development of Bitzer compressors for trans-critical applications with CO 2 is based on related compressor models and on two different housing sizes, the so called C3 and C4 housing. The types 4MTC and 4KTC are based on the compressor 4CC, whereas the models 4JTC, 4HTC and 4FTC are dissipated from the compressor type 4NCS. Both models, the 4CC and 4NCS, are designed for standard HFCs like e.g. R404A and R134a as well as for the HCFC R22. Some of the main differences are illuminated in the following. Oliver Javerschek, Dr. Günter Dittrich 1

2 2.1 Strength and safety The European standard EN defines the design criteria for commercial and industrial compressors. A safety factor of three can be applied in connection with type approved relief valves to the atmosphere. Taking a maximum operating pressure (MOP) of 75 bar on the low pressure (LP) side and 130 bar on the high pressure (HP) side as basis, the minimum burst pressures correspond to 225 bar and 390 bar respectively. In other words, the safety margin on LP and HP side is equal to 150 and 260 bar respectively. In order to full fill these requirements ductile materials like e.g. spherical cast have to be applied. Because of different casting processes, molts for grey cast iron cannot be applied. Concerning the type approved relief valves the standard EN 378 has to be applied. Taking the tolerances of the relief valves into account, the standard allows a pressure equal to the MOP multiplied by 1.1 until the maximum flow cross section of the relief valve is available. On basis of a MOP of 130 bar, the maximum pressure corresponds to 143 bar. But still 130 bar are valid for the safety against bursting. Further according to the EN 378, the high pressure switch has to be adjusted to a value that equals the MOP multiplied by 0.9, consequently 117 bar. 2.2 Reliability The load conditions for the drive gear of CO 2 compressors for trans-critical applications are determined by the characteristics of the natural substance: High pressures and pressure differences, high volumetric refrigeration capacities and remarkable high suction densities. In comparison to standard compressors like e.g. for the HFC R404A, the suction densities and volumetric refrigeration capacities lead to a compressor design with just small geometrical displacements for a similar cooling capacity at full load conditions at maximum ambient conditions. In this respect it has to be considered that the ratio of refrigerant mass flow in kg/h per m³/h geometrical displacement is much higher for CO 2 compressors. Focusing on the mechanical loads, the resulting gas forces are dependent on the cylinder bore and the maximum pressure difference. The resulting gas force for the 4NCS corresponds to kn on basis of a maximum pressure difference of 30 bar. The maximum pressure difference for the compressors for trans-critical applications corresponds to 110 bar (20 bar minimum suction - and 130 bar maximum discharge pressure). Therefore the resulting gas forces correspond to , and kn for different piston diameters for the models 4JTC, 4HTC and 4FTC. Consequently the drive gear contains strengthened connecting rods in the small and big eye to transmit the resulting forces. In order to prevent oval -and axial deformation of the piston pins under higher load conditions, they are designed to offer bigger bearing surfaces. A key factor of the design of CO 2 compressors for trans-critical applications is the ratio between stroke and bore. It has a significant influence on the efficiency and the reliability. With respect to the reliability of the compressors it is of utmost importance to achieve a safe dimensioning in terms of generously dimensioned piston pins, large seating for the piston pin inside the piston and large bearing surfaces inside of the connecting rods. Besides this the average piston velocity is of importance to minimize the wear. In order to achieve this, larger bores and consequently lower stroke-bore ratios are of advantage, because it offers simply more space than a design with longer strokes and smaller bores. The applied stroke-bore-ratios for the above mentioned compressor types is under square; this results in an average piston velocity of m/s and is reduced by 38.1 % in comparison to the compressor type 4NCS. While focusing on the reliability of semi-hermetic piston compressors in commercial applications, the kind of lubricant has a significant influence. Possible candidates for trans-critical applications with CO 2 are PVE, PAG and POE. The lubricants have to be evaluated with respect to compressor and system related issues. With focus to the compressor, hygroscopicity, di-electric strength, material compatibility, resulting viscosities, lubricity, thermal and chemical stability are of utmost importance. Besides the material compatibility, the miscibility with liquid CO 2 is relevant for the system design. In the following POE is considered as the lubricant. With respect to solubility and viscosity characteristics this choice is the most challenging for the compressor. But it offers advantages for the system design such as lower hygroscopicity than PAG, better miscibility and improved system oil circulation. The basic viscosity of the considered lubricant corresponds to 85 cst based on 40 C. An anti-wear package is applied with the oil. Oliver Javerschek, Dr. Günter Dittrich 2

3 The high gas solubility of CO 2 in POE can lead to strong de-gassing effects in connection with rapid pressure fluctuations inside the crank case. These effects occur inside the oil sump, the oil supply channels and on bearing surfaces. Strong de-gassing effects lead to boundary friction conditions, which require adequate measures to avoid severe wear of the drive gear. The pistons are based on standard alloy. During boundary friction conditions, standard pistons could not provide the requested reliability. Therefore the pistons are specially treated for improved running qualities. This offers low wear under extreme conditions, e.g. start-stop cycling with high liquid fraction in the suction gas. As mentioned above, lower stroke-bore-ratios offer the advantage of large bearing surfaces inside the connecting rods. The small eye contains a bearing out of high-performance material. It combines high reliability under high loads and good dry running capability during boundary friction conditions. Inside the big eye a multi layer bearing is applied which offers extremely good dry running capabilities. The same type of bearing is also applied with the main bearing and bearing inside the bearing cap. It was mentioned before that strong de-gassing effects can occur with fluctuating pressures inside the crank case of the compressors. The process of de-gassing is of course very much dependent on the amount of CO 2 that is solved in the lubricant and this is dependent on the suction pressure and oil sump temperature. While focussing on the application of variable speed drive (VSD), the resulting oil sump temperature is directly influenced by the resulting mass flows with different speeds. Considering a constant suction and discharge conditions, the oil sump temperature drops significantly with higher speeds. The opposite trend can be investigated concerning the oil carry over rates of the compressors. With higher speeds, and especially with speeds higher than 70 Hz, the oil carry over rate is increased. The development of the compressors considered these aspects; however they cannot be discussed within this paper. 2.3 Efficiency Characteristic values for the efficiency of compressors are the volumetric and isentropic efficiency. The volumetric efficiency is determined by the ratio of effective and geometrical volume flow. The major influences are additional superheating of the suction gas between the suction shut-off valve and the entrance into the cylinder due to motor losses, heat transfer from the cylinder surface and from discharge- to the suction side and valve malfunctions. Throttling, friction and motor losses show the biggest influence on the isentropic efficiency of compressors. The isentropic efficiency of the compressor can easily be calculated by the isentropic power input multiplied by the volumetric efficiency and divided by the measured power input at the terminals. Consequently this value includes all losses. The above mentioned values can also be described as total external efficiencies of a compressor. Indicated efficiencies determine internal losses. The indicated volumetric efficiency is defined as ratio of distance of the indicated volume to the swept volume. Therefore it describes the loss of useful suction volume due to back expansion and throttling losses. The indicated efficiency describes the ratio of isentropic power input and indicated power input. Energetic losses of the compression process due to leakages, throttling losses and heat transfer processes inside the cylinder can be detected. The indicated efficiencies will not be discussed in detail within this paper. However, a indicated p,v diagram will be used to describe the dynamic behaviour of the valves. In Chapter 3.2 the advantages of smaller stroke-bore-ratios for CO 2 compressors for trans-critical applications were highlighted. While discussing the pros and cons of a short stroke design, the efficiency also has to be considered. At first view, it seems that the disadvantages for the efficiency are dominant. A short stroke and consequently wider bores have the disadvantage of a bigger dead volume inside the cylinder. As well the tightening line for the piston rings is enlarged which leads to higher losses along the piston. Süss [1] proposed longer strokes in his study to reduce the leakages a long the piston. Another approach was considered for the design of the compressor types 4MTC to 4FTC in order to compensate the disadvantage with respect to the volumetric efficiency of a shorter stroke. The pistons are equipped with 3 piston rings. This solution ensures a reduction of the leakages along the piston. Besides this the triple piston ring solution provides a moderate pressure reduction with each stage and leads consequently to a reduced load on each piston ring. This increases the lifespan. Of course, the piston rings are also optimized for reduced friction to minimize losses with respect to the isentropic efficiency. Further, with respect to the volumetric efficiency, the dead volume of the compressor models is reduced to a minimum by observing a high safety. The advantages of wider bores with respect to the flow cross sections of the valve plate are obvious. The available space is used to design the flow cross sections and valve positions in an optimum way. The Bitzer compressor range for trans-critical applications with CO 2 is designed for VSD between 30 and 87 Hz. In order to minimize throttling losses not only at rated speed, but also with speeds between 70 and 87 Hz, flow cross sections must be designed accordingly. With higher mass flows, Oliver Javerschek, Dr. Günter Dittrich 3

4 the flow velocities are increased and the influence on the pressure drop is quadratic. Throttling losses lead to a lower indicated efficiency of the compressor. The available space is also utilized to apply an optimized valve design, which combines efficiency and high reliability. As a result, a delayed closing of the suction working and discharge working valves is prevented and leakages through the valves are minimized as well. Both malfunctions functions could reduce the indicated volumetric efficiency of the compressors but are avoided with the described design. The motor cooling of the compressor models 4MTC, 4KTC, 4JTC, 4HTC and 4FTC is realized by suction gas. By applying suction gas cooling, a balance point between necessary and excessive motor cooling has to be found. A suction bypass or low effects on motor cooling by the suction gas lead to the requirement of heat rejection through external oil cooling or cooling by ambient air. Too much motor cooling leads to very low oil sump temperatures caused by the extremely good cooling effect of CO 2 and to unnecessary high superheating of the suction gas with losses in terms of volumetric efficiency. Due to internal measures inside the motor compartment, the stream of suction gas is divided with the discussed compressor models. Internal channels ensure that a rated amount of suction gas bypasses the motor. Internal channels lead the suction gas directly to the suction chambers. A sufficient motor cooling is provided by the amount of suction gas that streams along the stator. Due to this measure, high volumetric efficiencies can be achieved and external measures for the motor and oil cooling do not have to be applied within the wide application range of the compressors. In order to minimize the heat transfer from discharge to the suction side at the cylinder heads, the assembly shows a new design. As shown in Figure 1, the suction and discharge chambers are separated. The reduced heat transfer results in further improved volumetric and isentropic efficiencies. Figure 1: Simplified drawing of the cylinder head Characteristics of the volumetric and isentropic efficiencies for the compressor model 4HTC-20K are provided in Figure 2. It shows a comparison of the first generation and the actual design. It defines the efficiencies on basis of a constant suction pressure of 28.0 bar and a suction gas superheat of 10 K; the pressure is equal to -8 C saturated suction temperature. Efficiencies are given for head pressures from 50 bar to 112 bar. 50 bar equals a condensing temperature of 15 C and 112 bar corresponds to a optimum discharge pressure for a gas cooler outlet temperature around 45 C. During the development of the compressor, the volumetric efficiency was improved significantly. On basis of a pressure ratio of 2.5, which equals a discharge pressure of 70 bar, an increase of 10 % could be achieved. From the beginning the cylinder heads featured separated suction and discharge chambers. With higher head pressures, which are typically for high discharge temperatures and a higher power input of the motor, the discharge temperatures show higher values as well. Consequently the suction gas between the suction shut-off valve and the suction chamber is influenced by higher waste heat of the motor and heat rejection from the heated cylinder walls and from the discharge side. The characteristics of the volumetric efficiencies show that the thermal load of the suction gas could be further reduced during the development. As well the working valves were optimized during the development. Besides the volumetric efficiency the isentropic efficiency could be improved significantly as stated in Figure 2. An increase of 12 % could be achieved for a pressure ratio of 2.5. Throttling and friction losses were reduced during the development. Taking a maximum gas cooler outlet temperature of 35 C and a discharge pressure of 90 bar into account, the pressure ratio corresponds to 3.2. Further considering a condensing temperature of 15 C, the total isentropic efficiency of the discussed compressor model is always clearly above This enables an efficient operation of the compressor throughout the seasons and leads to high values for the SEER. The fact that the isentropic Oliver Javerschek, Dr. Günter Dittrich 4

5 efficiency shows such a flat characteristic is typical for a piston compressor. The dead volume contains a certain amount of refrigerant. With the down motion of the piston, back expansion occurs and forces the piston down. Transmitted via the eccentric shaft, this fact reduces the necessary torque for the piston that moves up. Transcritical applications with CO 2 require efficient compressors under all operating conditions to reach similar or higher COPs than HFC baseline systems in moderate climate zones. Figure 2: Volumetric and isentropic efficiency of the compressor model 4HTC-20K Detailed investigations on the indicated work of the compressors were performed during the development and optimization of the compressors. The Figure 3 shows an example of an p,v diagram measured with the compressor model 4HTC-20K. Figure 3: Indicator diagram for VSD between 700 and 1900 rpm of the compressor model 4HTC-20K The function of the working valves was investigated in connection with VSD at constant suction and discharge conditions. All measurements are based on 26.5 bar suction and 90 bar high pressure, a suction gas superheat of 10 K and compressor speeds between 700 and 1900 rpm. Shortly after passing the lower dead centre, the piston moves upwards (1). The diagram demonstrates that the compression lines (1) to (2) with various compressor speeds are almost identical; therefore it demonstrates that the closing behaviour of the suction valves is good under various mass flow rates. The exhaust stroke is shown between (2) and (3). The various characteristics show a certain deviation concerning the over compression. Self evident, certain pressure differences between discharge and high pressure are necessary for unloading the cylinder and discharging into the high pressure Oliver Javerschek, Dr. Günter Dittrich 5

6 chamber of the cylinder head. The measured pressure differences are in the range of 3 to 11 bar between 700 and 1900 rpm. At the upper dead centre of the piston, the discharge valve closes and as soon as the piston starts to move down again, back-expansion takes place between (3) and (4). The remaining gas inside the dead volume of the cylinder expands back to suction pressure (4). The almost identical characteristics prove a good closing behaviour of the discharge valves. The pressure differences between the opening pressures of the suction valves and the suction pressure are stated between (4) and (1). The differences are in the range of 2 to 6 bar between 700 and 1900 rpm. The indicator diagram highlights that the working valves are designed to an optimum for the application of VSD. A Malfunction like a delayed closing of the valves does not occur. While assessing the discharge- and suction characteristics in Figure 3, it has to be considered that working valves have a certain inertia. The mechanical inertia and the spring force of the working valve result into a resistance that acts against the opening force. With higher compressor speeds and consequently higher mass flows and flow velocities, the resistance is increased as well. A higher force leads to higher pressure differences for the opening and closing of the working valves. An additional effect shows the oil on the valve seats that causes an adhesive bonding effect. Besides investigations for different compressor speeds, tests were carried out for various head pressures as well. However, within the scope of this paper all aspects of the development and the optimization of the compressors for trans-critical applications cannot be illuminated. 3 Various systems for commercial applications Very often all CO 2 systems for MT and LT are applied as a cascade or externally compounded booster system in commercial applications. Mostly these systems are provided with a flash gas bypass (FGB) in the MT stage. These systems can be considered as state of the art while thinking globally (in terms of number of installed units), CO 2 is still far away of being a standard solution. A commercial all CO 2 cascade system for MT and LT, including an FGB, was commissioned for the first time in November 2004 by Linde. A detailed description of such a system design can be found in various publications, e.g. Haaf et al. [2]. With respect to a better SEER, cost reduction and simplified system design, booster systems become more and more the preferred option. Booster systems have a common refrigerant circuit and are characterised by direct exchange of refrigerant, oil and heat energy. Naturally the booster systems do not apply a cascade heat exchanger. This paper discusses the question, whether there are other alternatives that can be applied in the field. 3.1 Booster system for MT and LT with external flash gas cooling A booster system with external flash gas cooling is shown in Figure 4. Down stream the gas cooler the refrigerant flows through the valve HP control valve and expands to a receiver. Figure 4: Booster system with external flash gas cooling Liquid and flash gas becomes separated. The gas cooler outlet conditions (pressure and temperature) determine the ratio of flash gas inside the receiver for a defined medium pressure level. Contrary to a system with flash gas bypass, the amount of flash gas streams through a heat exchanger by the principle of natural circulation and Oliver Javerschek, Dr. Günter Dittrich 6

7 condenses back into the receiver. The flash gas is cooled by an additional refrigeration circuit with evaporation on the cold side of the heat exchanger. Due to this, the complete mass flow of the booster system is available as liquid down stream the receiver for the MT and LT evaporators. Despite of the changes concerning the procedures around the flash gas, the system is equal to a standard booster system. The disadvantage is the higher complexity. On the other hand, the additional cooling unit can be applied useful during the initial commissioning and during stand still periods. With focus on the COP, the power input of the condensing unit shows an additional energetically effort. While applying a refrigerant for this unit which is efficient at high ambient conditions, such a hybrid solution results into a better COP than in comparison the standard booster system with flash gas bypass. 3.2 Booster system for MT and LT with internal sub cooler and high and medium pressure vessel Another alternative is discussed in Figure 5. The complete mass flow of the system expands into a so called high pressure receiver. Again the gas cooler outlet conditions (pressure and temperature) determine the condition inside the receiver. In case of no separating of liquid and flash gas, the pressure inside the receiver corresponds to pressure at bulb temperature. The pressure reduction reduces the mechanical load on the sub cooler downstream the first receiver. Saturated liquid at bulb temperature streams out of the receiver and is sub cooled in counter flow by refrigerant at medium temperature pressure level in an ideal case. The liquid sub cooling is dependent on the amount of refrigerant, which is used for the internal cooling effect. A possible solution would be to apply the same refrigerant mass flow as generated flash gas for a defined pressure level in case of a process with FGB. Figure 5: Booster system with internal sub cooler and high and medium pressure vessel Due to this fact, pressure levels below 40 bar could be adjusted as in systems which apply the standard flash gas removal. Downstream the sub cooler, the refrigerant expands again into a receiver, the so called medium pressure receiver. Without flash gas removal the pressure inside the receiver again equals the pressure at bulb temperature. Downstream the second receiver the refrigerant mass flow is divided into the MT, LT and sub cooler mass flow. Upstream the MT compressor stage, the complete mass low is generated again and taken in by the MT compressors. 3.3 System comparison A comparison of the before described systems is based on the following assumptions: Evaporating temperatures of -10 C and -35 C for MT and LT, discharge pressure of 90 bar for the MT stage, 35 C gas cooler outlet temperature, 20 K superheat at the compressor inlets, 8 K useful superheat for MT and LT evaporators and an equal geometrical displacement for the MT and LT stage. An externally compounded two stage booster system with FGB in the MT stage is considered as a benchmark. Further it is assumed that the medium pressure equals Oliver Javerschek, Dr. Günter Dittrich 7

8 35 bar. For the system with external flash gas cooling it is assumed that 0.5 kw power input results in 1 kw cooling capacity. A reciprocal value of the COP in the range of 0.5 is a modest value on basis of an evaporating temperature of -10 C (10 K temperature difference to liquid temperature inside the medium pressure receiver) and a condensing temperature of 45 C based on an ambient temperature of 35 C. By applying the natural refrigerant propane or the HFC R134a, values lower than 0.43 kw power input per 1 kw cooling capacity could be achieved under the mentioned conditions. For the FGB process it has to be considered, that the expansion process of flash gas back to MT evaporating pressure ends inside the two-phase area of CO 2. The generated amount of liquid has to be evaporated, e.g. inside a heat exchanger, before it enters the suction ports of the compressors. This is to avoid wet operation. The calculation of the FGB process does include a liquid to flash gas heat exchanger which creates a certain amount of liquid sub cooling. A useful superheat of 8 K is considered on the cold side of the heat exchanger. The same value for a useful superheat is applied for the sub cooler of the system 3.2. A reduced amount of flash gas inside evaporators due to the application of flash gas removal offers the potential for an increase of the heat transfer coefficient and reduced pressure drops on the CO 2 side [3]. These effects are not considered for this comparison. While focusing on the theoretical COP of the system with external flash gas cooling, the value is improved. The comparison shows an increase in MT and LT cooling capacity of 67.0 % and an increase in power input of 51.1 % due to the flash gas cooling unit. This results in an increased COP of 10.5 %. Taking the described gas cooler outlet conditions and medium pressure level into account, the ratio of liquid and flash gas inside the receiver of the baseline system corresponds to 56.4 and 43.6 % respectively. By considering a mass flow through the sub cooler of the system 3.2 that equals the amount of flash gas in a system with FGB, identical medium pressures inside the medium pressure vessel are achieved. Consequently the available enthalpies for the MT and LT evaporators are equal. On basis of the defined suction - and discharge conditions, the power input is the same. Consequently the COP of the system 3.2 corresponds to the value of the baseline system. 4 Conclusions Compressors for trans-critical applications with CO 2 require demanding techniques with the respect to strength and safety, reliability and efficiency. By applying innovative ideas and new technologies, the mentioned compressor models for trans-critical CO 2 applications have reached an advanced level of engineering. The compressors feature no external piping, suction gas cooled motors without the requirement for oil cooling, offer high efficiencies and very favourable operating characteristics with low vibrations and pulsation due to 4 cylinder design. From the application engineering point of view the compressors are designed to withstand challenging operating conditions, like e.g. wet operation and boundary friction conditions. The compressors are designed for VSD in a wide range from 30 up to 87 Hz and combine high reliability with superior efficiencies. Concerning alternative system designs for trans-critical CO 2 applications, two new systems were introduced. The priority of utmost importance for CO 2 systems is: Increasing the efficiencies by reducing the complexities. The benchmark system is an externally compounded two stage booster system, which applies an FGB in the MT stage. The complexity is on an acceptable level and in comparison to R404A systems, the efficiencies are promising in moderate climate zones. Therefore, only the proposed system with external flash gas cooling shows an alternative. For the case that a factory assembled hybrid solution is accepted, the system with external flash gas cooling could be applied. The complexity is different but on a similar level and the efficiency is improved. 5 Nomenclature COP coefficient of performance ( ) FGB flash gas bypass GWP global warming potential HCFC hydroclorinefluorcarbons HFC hydrofluorcarbons HP high pressure LT low temperature MOP maximum operating pressure MT medium temperature PAG polyalkylenglycol Oliver Javerschek, Dr. Günter Dittrich 8

9 POE PVE SEER TEWI VSD polyolester polyvinylether seasonal energy efficiency rating total equivalent warming impact variable speed drive 6 References and Bibliography 1. Süß J. 1998, Untersuchungen zur Konstruktion moderner Verdichter für Kohlendioxid als Kältemittel, DKV Forschungsbericht Nr. 59: 2. Haaf S., Heinbokel B., Gernemann A. 2005, Erste CO 2 -Kälteanlage für Normal- und Tiefkühlung in einem Schweizer Hypermarkt, Die Kälte & Klimatechnik 2/2005: Ebel S., Hrnjak P., 2004, Flash gas bypass for improving the performance of trans-critical R744 systems that use microchannel evaporators, International Journal of Refrigeration 27 (2004): Oliver Javerschek, Dr. Günter Dittrich 9