Experimental analysis of R744 parallel compression cycle

Applied Energy 135 (2014) 274–285

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Experimental analysis of R744 parallel compression cycle Andrea Chesi a,1, Fabio Esposito a,1, Giovanni Ferrara a,1, Lorenzo Ferrari b,⇑ a b

Dept. of Industrial Engineering, University of Florence, Via di Santa Marta, 3, 50139 Florence, Italy CNR-ICCOM, National Council of Research of Italy, Via Madonna del Piano 10, 50019 Sesto Fiorentino, FI, Italy

h i g h l i g h t s

g r a p h i c a l a b s t r a c t

 Theoretical analysis of the parallel

compression cycle in flash tank configuration.  Analysis of the influence of flash tank separation capacity on system performance.  Coupled experimental and numerical analysis.  Main detrimental factors detected.  Best COP improvement of 10% and limited by ineffective separation.

a r t i c l e

i n f o

Article history: Received 2 December 2013 Received in revised form 2 July 2014 Accepted 25 August 2014

Keywords: Experimentation R744 Parallel compression Flash tank Separation effectiveness

a b s t r a c t A diffuse interest on cycle modifications for R744 applications that allow to achieve better CO2 cycles performance can be inferred from the literature. The parallel compression cycle in flash tank configuration seems to be an interesting solution and the authors conducted an experimental campaign to estimate the critical parameters that influence its performance in an actual cycle. Previous to that a thermodynamic analysis was lead to investigate the expectable performance of the cycle and the influence of key parameters such as compressors volumetric flow ratio and separator efficiency. The thermodynamic model designed also allowed the definition of the limit conditions for an optimal operability of the cycle of interest. The experiments showed that the theoretically reachable improvements in terms of refrigerating capacity and coefficient of performance are threatened by several phenomena which may occur in a real system. These detrimental effects are described and commented in the paper. The paper also includes experimental data obtained during the experimental activity for operating conditions typical of commercial refrigeration applications. Ó 2014 Elsevier Ltd. All rights reserved.

1. Introduction ⇑ Corresponding author. Tel.: +39 055 5225 218; fax: +39 055 5225 203. E-mail addresses: [email protected]fi.it (A. Chesi), [email protected]fi.it (F. Esposito), [email protected]fi.it (G. Ferrara), [email protected] (L. Ferrari). 1 Tel.: +39 055 4796 570; fax: +39 055 4796 342. http://dx.doi.org/10.1016/j.apenergy.2014.08.087 0306-2619/Ó 2014 Elsevier Ltd. All rights reserved.

Stratospheric ozone depletion and the risk of global warming is pushing international environment associations and national administrations towards the adoption of refrigerants different from the present widely implemented synthetic ones, such as

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275

Nomenclature

Latin c h m P Q T V W x Greek /

g k

q

specific heat, kJ kg–1 K–1 specific enthalpy, kJ kg–1 mass flow rate, kg s–1 pressure, MPa capacity, heat flow rate, kW temperature, °C swept volume, m3 s–1 power, kW quality, –

separator efficiency, – energy efficiency, – volumetric efficiency, kg m–3 density

Subscripts BFT bottom of the flash tank cmp1;1 compressor 1 cmp2;10 compressor 2

HFCs. Natural fluids are becoming an attractive field of research in refrigeration thanks to their very low environmental impact indexes. Among the most appealing natural fluids stands carbon dioxide (R744) which is featured by a negligible contribution to greenhouse effect and no ozone depletion potential, furthermore it is considered one of the safest alternatives for what concerns toxicity and flammability [1]. The major drawback of the use of carbon dioxide as a refrigerant is its low critical temperature (31.1 °C) which makes it impossible to avoid transcritical cycles when rejecting heat at ambient temperature typical of the summer season of most of the countries in the temperate belt. This kind of cycle is afflicted by higher exergy losses during both gas cooling phase and the throttling. The first is caused by the transcritical cooling which is not isotherm, the latter and most significant [2] by the higher pressure drops required with CO2. Another disadvantage is the high pressure (up to 14 MPa) which all the components of the refrigeration system must bear. In general the single stage compression with single throttling cycle, which is the simplest among those evolving CO2, is not capable of performance up to those of synthetic refrigerants. Thus, whereas in the past the research focused on the fluid itself and its properties, now, once it is assumed to use carbon dioxide as refrigerant, research focus must be on the thermodynamic cycle exploited and the technology implemented for its fulfilment in order to achieve commercially appealing performance. Several solutions presented in the literature implement cycle modifications to achieve better performance, e.g. multi-stage compression, cycles with internal heat exchangers, economized cycles etc. Among the most interesting solutions in terms of coefficient of performance (COP) and cooling capacity increase there is the parallel compression with flash vapour suction. The main idea behind the concept of parallel compression cycle is to reduce the throttling losses [3]. In the case of flash tank configuration this reduction can be obtained splitting the throttling phase between two different valves, separated by a mid-pressure receiver which works as a separator between the liquid and the vapour phase. Once the separation took place it is possible to

diss ess eV FTI gc in int lim;Lim max out;o s suc V 5 6 7

dissipator equivalent single stage evaporator flash tank inlet gas cooler inlet intermediate limit maximum outlet isentropic suction volumetric flash tank inlet flash tank outlet towards auxiliary compressor flash tank outlet towards evaporator

Abbreviations COP coefficient of performance ESS equivalent single stage IHX internal heat exchanger

remove the flash gas from the fluid by means of an auxiliary compressor: this prevents the flash vapour from entering the evaporator, where it would not provide any useful effect, thus being sequentially compressed by means of wasted work. In fact extracting the vapour from the flash gas allows to compress it with a lower compression ratio. Another remarkable positive effect is that the cooling capacity is increased because the fraction of carbon dioxide entering the evaporator has a quality nearer to zero than it has without the use of flash tank. The aim of the activities presented in this paper is to investigate, both theoretically and experimentally the parallel compression cycle in flash tank configuration, in order to estimate how the thermodynamic parameters affect system performance, the intrinsic limits on temperature and pressure ranges of this particular layout and the experimental critical points to overcome. A thermodynamic analysis model was implemented and the theoretical performance increments, both in terms of COP and cooling capacity, were estimated for different evaporation pressures, compressor discharge pressures and gas cooler outlet temperatures. It was also investigated the influence on performance of the compressors volumetric flow ratio and flash tank separation capacity. Furthermore the limits for the operating conditions of the parallel compression cycle were defined in respect to the parameters just mentioned. The parallel compression cycle in flash tank configuration was tested experimentally at the multipurpose test rig for CO2 refrigerating cycles analysis of the Department of Industrial Engineering of the University of Florence [4] with a fixed value of compressors volumetric flow ratio. The tests carried out investigated different ranges of evaporation pressures, for each one the gas cooler exit temperatures and discharge pressure were varied. The data collected allowed the evaluation of the critical parameters for the parallel compression and their effect on cycle performance. In addition it was possible to analyse the limits of the range of maximum pressure at which this configuration of parallel compression can run with a fixed displacement and revolution speed auxiliary compressor while having

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a pure liquid phase at the exit of the separator heading to the main compressor. Moreover the critical physical processes that ensure or endanger the expected performance increase were evaluated thanks to the data acquired. The critical aspects regarding the experimentation of the parallel compression in flash tank configuration will be depicted. 2. State of the art On the contrary of what happened with synthetic refrigerants, with whom the research focus was on the fluid properties, the implementation of carbon dioxide determines a different direction for researchers. Indeed more elaborated thermodynamic cycles and advanced component technology are required. The simple single-stage compression single-throttling cycle has been widely investigated by many authors in [5,6]. Other possible cycle configurations are: two-stage compression with intercooler [7], cycles involving an internal heat exchanger where the flows coming out from the evaporator and the gas cooler establish a heat exchange [8–11] and economized two-stage compression cycles, where part of the mass flow is spilled from the mid-pressure tank, throttled and then used to lower the temperature of the gas flowing towards the second throttling valve [12]. Many authors presented various works, both theoretical and experimental about these base-cycle modifications with results that underline the benefits that can be achieved in specific working conditions. One of the most promising cycle layouts in terms of expected performance increase is the parallel compression. An example is provided by the pressure-specific enthalpy diagram in Fig. 1. The cycle delineated by the continuous black line relates to the parallel compression in flash tank configuration whereas the simple single-stage compression cycle is partially represented with a broken line and delimitated by points 1-2-4-12. It is possible to notice the usual gas cooling (3-4), the main compression (1-2), the flash gas compression (10-11), the first throttling (4-5), the liquid–vapour separation (5-6-7), the second throttling (6-8), the carbon dioxide evaporation (8-9) and the two overheating (7-10 for the auxiliary compressor suction line, 9-1 for the evaporator and main compressor suction line). There are different layouts that rely on the parallel compression concept which have been discussed in the literature. In [13] Bell covers with a theoretical analysis a generally economized cycle, providing the expected performance in terms of COP and cooling capacity for different gas cooler outlet temperatures and economizer pressure, depending also on the compressor swept volume ratio, concluding that this ratio needs to change if an optimization for a wide range of gas cooler outlet temperature is desired. Agrawal et al. [14], focuses on three cycles layout: two-stages compression with flash gas bypass, two-stages cycle with flash gas intercooling and two-stages cycle with compression

Fig. 1. P–h diagram for the parallel compression cycle in flash tank configuration.

intercooling (which is not a parallel compression cycle); the authors conclude that the simultaneous optimization of the maximum and intermediate pressure, for given evaporation and rejection temperatures, leads to a maximum value of the COP and that the highest value is reached by the flash gas bypass layout. Another critical point that is underlined in the paper is the possibility to decrease the optimum discharge pressure with this particular layout compared to the pressure required by the single-stage cycle. Sarkar and Agrawal [15] investigate theoretically three different configurations and identifies the parallel compression economized cycle as an interesting solution to improve thermodynamic performance of CO2 cooling devices especially for lower temperature applications, achieving a 47% COP improvement over the basis cycle in the chosen ranges of operating conditions. Cecchinato et al. [16] confirm how the most elaborated cycles such as the Double-Throttling, Double Compression in both Open Flash Tank and Split Cycle configuration present the greatest improvement over the single stage basic cycle. In the theoretical analysis carried out it is observed that the energy efficiency for this particular layout can reach improvements in the order of 70% for the heaviest operating conditions (such as 30 °C evaporation temperature and 35 °C gas cooler outlet temperature). Da Ros [17] investigates the optimization of parallel compression with flash tank, obtained by means of compressors discharge pressure and displacement ratio tuning, providing values for COP and discharge pressure of both compressors in a few operating conditions. Cecchinato et al. [3] provided an important theoretical analysis of the cycle and its key parameters. From the work it could be observed that the ratio between the compressor displacements has been confirmed as an important influencing factor on the performance. Moreover, given the maximum pressure of the cycle and the environmental conditions, there is a relation between the intermediate pressure and the displacement ratio. The COP is function of both maximum and intermediate pressure and it is not possible to optimize the performance over a wide range of temperatures without using compressors capable of varying their swept volume; at the end of the work an experimental test rig is proposed for supplementary analysis. According to the literature suggestions, a parallel compression cycle with flash tank will be investigated both theoretically and experimentally in the present paper in order to estimate its theoretical and actual performance, its operating limits and the impact of several factors on the performance itself. 3. Thermodynamic analysis A theoretical analysis was carried out to assess what performance to expect from the layout of interest, the critical parameters influencing cycle performance, the range of temperatures and discharge pressures that the parallel compression in flash tank configuration allows. The numerical model used to simulate parallel compression cycles has its basis in the libraries provided by Span and Wagner [18] for the thermo-physical properties of the refrigerant and it is solved by means of a commercial software. Some hypotheses were assumed:  Steady-state conditions.  No pressure losses, excluded those intended (only throttling phases in the ideal cycle).  No heat exchanges with the environment, excluded those intended (heat exchangers and suction lines of the compressors in the ideal case).  Compressor efficiency and volumetric efficiencies are correlated to the compression ratio on the basis of previous analysis on compressors themselves.

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 Maximum versatility on superheat at the compressor suction lines and inside the evaporator. The model inputs may vary depending on what the model seeks out, but in general are:  Evaporation pressure.  Compressor discharge pressure.  Gas cooler outlet temperature.  Separator efficiency value (described in this section).  Compressors volumetric flow ratio. From these parameters it is possible to determine all the thermodynamic states involved in the cycle and its global performance. Parallel compression adoption allows one to get an additional degree of freedom for the achievement of optimal performance which is the intermediate pressure between the discharge/gas cooler and the suction/evaporator ones. In our case this pressure is the one acting inside the flash tank and from which the auxiliary compressor is sucking CO2. The compressors volumetric flow ratio, defined by Eq. (1), is a key parameter for the analysis of parallel compression cycles and it was introduced in [3] in a slightly different fashion as ‘‘compressor displacement ratio’’.

RV ¼

V_ aux _V main

ð1Þ

One of the first analysis carried out in this study pointed to assess the influence of the compressors volumetric flow ratio (and consequently of the pressure inside the flash tank) on the COP. The results for an evaporation pressure of 2.25 MPa (corresponding to a temperature of 15 °C) and a gas cooler exit temperature of 35 °C are illustrated in Figs. 2 and 3. For each couple of evaporation pressure and gas cooler outlet temperature it is possible to determine a discharge pressure and intermediate pressure couple that optimize the COP. It is important to notice though that at a fixed discharge pressure, assuming 8.5 MPa, while the intermediate pressure varies for example from 4.5 MPa to 7.2 MPa the volumetric flow ratio between the compressor varies respectively from 0.51 to 0.18. This means that to obtain this level of optimization both the compressors in the system must be capable of varying their volumetric flow over a wide range. At an increase of compressors volumetric flow ratio, in the theoretical case of complete vapour–liquid separation, follows a decrease of the intermediate pressure inside the flash tank. There are three possible ways to modify the compressors volumetric flow ratio:

Fig. 3. Influence of compressors volumetric flow ratio on COP.

 Choosing different compressors with the desired displacement, this is something that can be done only during the design of the system and cannot be modified in real time.  Varying the speed of revolution of one or both the compressors by means of one or two inverters, this solution provides a more versatile system but at higher costs and complexity.  A combination of the two above. In the experimental analysis the compressors volumetric flow ratio was kept constant for two reasons: the first because in commercial applications, it is rare to find a system working with variable volumetric flow ratios and the second for this allowed the authors to better understand the behaviour of the flash tank. The physical behaviour of the flash tank is described hereafter. The intermediate pressure of the circuit is a function of: the gas cooler outlet conditions (T4, p4 if we consider the numeration of Fig. 1), the volumetric flow of both compressors, their volumetric efficiency and suction density. During the experimental campaign disserted in Section 4 the gas cooler outlet temperature and pressure and the volumetric flows of the compressors were measured directly whereas the volumetric efficiencies and the suction densities were calculated: from correlations in the first case and from the measurement of temperature and pressure at the compressors inlets in the latter. Because the ducts from the flash tank to the compressors are not branched the mass flow at the main compressor must be the same extracted from the bottom of the flash tank and the mass flow elaborated by the auxiliary one must equal the mass flow extracted from the upper tap of the vessel. Eqs. (2) and (3) derive from this concept. The subscripts of the variables are in compliance to the numeration of Fig. 1.

V_ main q1 kv ;cmp1 ¼ V_ 6 qL6

ð2Þ

V_ aux q10 kv ;cmp2 ¼ V_ 7 qV7

ð3Þ

In Eqs. (2) and (3) there are two implicit assumptions: from the bottom tap of the flash tank exits only liquid and from the upper tap flows only vapour. This situation describes what happens if the system is working in ideal, or ‘‘optimal’’, conditions. From Eqs. (2) and (3) the volumetric flows from the upper and bottom tap of the flash tank can be determined. In steady-state conditions the mass flows entering and exiting the vessel must be balanced, Eq. (4).

_ 7þm _6¼m _5 m Fig. 2. Coefficient of performance variability over intermediate pressure and discharge pressure change.

ð4Þ

Considered the hypothesis introduced regarding the quality of the flow that comes out from the flash tank Eqs. (5) and (6) can be derived.

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_6¼m _ 5 ð1  x5 Þ m

ð5Þ

_7¼m _ 5 x5 m

ð6Þ

The ratio between the volumetric flows exiting the flash tank can be obtained from Eqs. (2), (3), (5) and (6), and it is expressed in Eq. (7).

V_ 7 x5 qL6 ¼ _V 6 ð1  x5 Þ qV7

ð7Þ

Anyway, in real cases, it is more common to know the volumetric flow of the compressors, thus Eq. (7) can be modified by means of Eqs. (2) and (3) so to take into account the compressors volumetric flow ratio, see Eq. (8).

RV ¼

V_ aux x5 q1 kv ;cmp1 ¼ _V main ð1  x5 Þ q10 kv ;cmp2

ð8Þ

Eq. (8) allows one to understand the behaviour of the flash tank. Let one consider a given (and at least for now, constant) value of RV to which corresponds a volumetric flow for each of the two compressors installed (constant as well). Starting from a given stable condition for the gas cooler outlet the fluid is laminated keeping the same specific enthalpy during the transformation. As it enters the flash tank it assumes different quality values depending on the pressure set inside the tank itself. For the given conditions (RV, T5, p5 and k and q for both compressors) if it exists it is unique the intermediate pressure which ensures both of the following conditions:  Only liquid exiting from the bottom tap of the separator and only vapour from the upper tap;  Steady-state conditions inside the vessel. If the steady-state condition is not respected then the mass flow balance in Eq. (4) it is not verified and the vessel is either emptying or getting full of liquid. If the conditions exposed cannot be met then there are two possibilities:  The compressors volumetric flow ratio can be varied so that the mass flow balance can be verified again.  It must be accepted a fraction of liquid exiting from the upper tap or a vapour fraction from the bottom one. In parallel compression systems the auxiliary compressor is usually smaller than the main one. Therefore it can happen that for relatively small pressures at the exit of the gas cooler and high gas cooler outlet temperatures the amount of vapour at the vessel inlet might be so high that the auxiliary compressor cannot remove all of it from the tank. The intermediate pressure may rise so that the increased density allows more vapour to be extracted for the same volumetric flow. With the pressure also the density at the inlet rises but not with the same rate: at some point the mass flow balance inside the vessel is no longer verified and a limit condition is reached. Thus if RV has a fixed value then it is not always possible to achieve only liquid conditions at the bottom tap of the tank. It is possible to define a parameter which can help the evaluation of the separator capacity to effectively reach the limit conditions, in other words, the minimum value of vapour extracted from the bottom tap. In this paper this parameter will be simply called separator efficiency defined in Eq. (9):

u¼1

hBFT  hLim hFTI  hLim

conditions) flowing out from the bottom tap of the vessel and hFTI is the specific enthalpy of the CO2 at the entrance of the flash tank. This parameter assumes the value 0 when no separation is taking place inside the flash tank and value 1 when the separation is complete. In the latter case, from the liquid side tap exits liquid with possibly the smallest amount of vapour which is directed towards a second throttling valve whereas from the vapour side tap exits only vapour that reaches the auxiliary compressor. On the other hand when the separation is not complete (u < 1) both from the upper and bottom taps of the flash tank might exit a liquid–vapour mixture with different quality. Actually during the experiments it was verified that, thanks to the geometry of the tap, from the upper tap always exits only vapour. Once both operating conditions and compressors volumetric flow ratio are defined, the separator efficiency and the intermediate pressure are related, as one decreases the other does the same. The diagram in Fig. 4 shows the influence of the separator efficiency over COP at fixed maximum and evaporation pressures for three different values of gas cooler exit temperatures. In the theoretical model, Eq. (9) allows the resolution of the physical problem, determining the intermediate pressure inside the flash tank together with the compressors volumetric flow ratio once that the value of hLim is defined for the conditions tested. The lack of a complete separation within the flash tank is very detrimental for performance, in some conditions it is even possible to lose completely the positive effect of the parallel compression. The tested configuration should be capable of great improvements on cycle performance. The diagram in Fig. 5 shows the COP per cent increase as a function of the discharge pressure and gas cooler outlet temperature. For the first thermodynamic analysis of the cycle performance some boundary conditions were fixed. The superheating within the evaporator from the saturation temperature and the superheating long both compressors suction lines are considered constant at a value of 5 °C; the pressure loss along the pipelines are null; the compressors volumetric flow ratio is that of the experimental analysis presented in this paper (RV = 0.28) and the separator efficiency is considered as a constant as well at value 1. The diagrams in Figs. 5 and 6 illustrate the performance improvements that can be expected with the boundary conditions just described. The dotted lines show the resulting performance improvements for each gas cooler exit temperature. The black lines connecting every series, denominated ‘‘Locus of maximum COP’’ and ‘‘Locus of maximum cooling capacity’’ respectively in Figs. 5 and 6, links together the performance improvements points where the maximum COP is scored by the parallel compression. Thus, even if the simple performance increase would suggest that the best improvement for the Tgc,o = 35 °C is at 8.5 MPa discharge pressure, the locus of maximum COP lines clearly shows that the best COP will be

ð9Þ

hBFT is the specific enthalpy of the actual flow exiting from the bottom of the flash tank, hLim is the specific enthalpy of carbon dioxide with the lowest possible amount of vapour (at the actual

Fig. 4. Influence of the separator efficiency on COP.

A. Chesi et al. / Applied Energy 135 (2014) 274–285

Fig. 5. COP percentage increase as a function of gas cooler outlet temperature and compressor discharge pressure for the given operating conditions.

279

Fig. 7. Operability limits of parallel compression cycle at different evaporation temperatures at RV = 0.28 and only liquid exiting from the bottom port of the flash tank.

Fig. 6. Cooling capacity percentage increase as a function of gas cooler outlet temperature and compressor discharge pressure for the given operating conditions.

achieved at 9 MPa instead. Another result that can be inferred from the diagrams is that the parallel compression is most valuable at harsh environmental conditions such as high gas cooler outlet temperature. The results of the theoretical model turned out to be in agreement with those of previous studies such that reported in [19] where several cycle modification to recover throttling losses in CO2 systems were proposed. At the chosen volumetric flow ratio and having considered only liquid exiting from the bottom tap of the flash tank, not all the curves span the same range of discharge pressure. This fact guides the analysis to a further step, where the thermodynamic model was set to estimate the limits of optimal operability of the parallel compression, depending on the evaporation pressure and the two factors RV and u. For ‘‘optimal’’ it is meant that the limit condition inside the vessel is the one with only liquid flowing from the bottom tap, which leads to the smallest amount of vapour to be compressed from the evaporation pressure instead than from the intermediate one. The diagrams in Figs. 7 and 8 show how the range of discharge pressure at which the parallel compression can be obtained in a steady condition varies. For a given set of evaporation pressure, compressors volumetric flow ratio and separator efficiency the continuous lines delimit the area where the parallel compression can work in the optimal way. In both cases the range of investigation is between 8 MPa and 13 MPa, because these are considered as reasonable limits for real applications. The same reason explains why the upper and lower limit for gas cooler outlet temperature are 20 °C and 45 °C. Higher or lower rejection temperature will not be investigated.

Fig. 8. Operability limits of parallel compression cycle with different compressors volumetric flow ratios at Tev = 5 °C and only liquid exiting from the bottom port of the flash tank.

The upper limit for the discharge pressure exists because for higher values the intermediate pressure should be lower than the evaporation pressure; on the other hand the lower limit exists because for lower values of the discharge pressure, the auxiliary compressor should extract more flow than the amount allowed by the RV ratio. Fig. 7 shows that at higher evaporation pressures the limits narrows down. The influence of displacement ratio is shown in the diagram in Fig. 8 that compares, at a fixed evaporation pressure, three different system conditions. Bigger auxiliary compressors allow one to achieve lower minimum discharge pressure at the cost of a maximum discharge pressure reduction and, more important, a tighter range of gas cooler exit temperatures. 4. Experimental activity 4.1. Experimental setup In order to analyse the performance of the parallel compression in flash tank configuration, a dedicated experimental set up was implemented on the multipurpose test rig of the Department of Industrial Engineering of the University of Florence [4]. The test rig was extensively tested and validated in previous activities not reported here for sake of brevity. A detailed analysis of the test

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rig performance can be found in [4]. The test facility allows the evaluation of several cycles performance thanks to the number of different components that can be implemented, nonetheless it allows the performance evaluation of external components which can bypass the original ones by means of three ways valves. The test rig is provided with a dedicated thermal dissipation system, measurement and acquisition equipment. In Fig. 9 it is illustrated a simplified layout of the experimental setup adopted. In order to guarantee an optimal lubrication of the compressors an oil separator vessel is installed in the circuit immediately downstream of the high-pressure manifold. The oil is gathered at the bottom of the vessel and from there is injected inside the compressors. The ‘‘T’’, ‘‘P’’ and ‘‘M’’ letters identify the temperature, pressure and mass flow acquisition sections respectively; moreover, for both the compressors the power absorption is monitored and acquired. The mass flows are acquired by means of two Coriolis flow meters; a detailed list of the sensors used to collect the data during the tests is presented in Table 1. The raw data obtained permitted the calculation of: compressor volumetric efficiency, Eqs. (10) and (11) and compressor global efficiency, Eqs. (12) and (13).

kv ;cmp1 ¼

kv ;cmp2 ¼

mCO2 ;cmp1

qcmp1;in V cmp1 mCO2 ;cmp2

qcmp2;in V cmp2

Accuracy

WIKA-S1 pressure transducers WIKA-S1 pressure transducers WIKA-S1 pressure transducers Type T thermocouples Yokogawa Rotamass RCCT-34 Siemens Massflow MASS-2100 HT Power Meter GSC57

bar bar bar °C kg h1 kg h1 kW

60 120 160 400 5000 1000 –

0.25% FSO 0.25% FSO 0.25% FSO 0.5 0.55% MV 0.10% MV 1% MV

COP ¼

Q ev W cmp1 þ W cmp2

ð16Þ

Q ess ¼ mCO2 ;cmp1 bhev;out  hgc;out c

gcmp2 ¼

mCO2 ;cmp2 ðhcmp2;out;s  hcmp2;in Þ W cmp2

ð13Þ

The data acquired allowed also the calculation of cycle performance, through gas cooler capacity (Eq. (14)), cooling capacity (Eq. (15)) and COP (Eq. (16)):

gev

Range

ð15Þ

ð11Þ

ð12Þ

mev;diss cev;diss bT ev;diss;in  T ev;diss;out c

Units

ð14Þ

ð10Þ

mCO2 ;cmp1 ðhcmp1;out;s  hcmp1;in Þ W cmp1

Q ev ¼

Sensor

Not being able to estimate the quality of the carbon dioxide at the evaporator inlet, the cooling capacity has been estimated by means of the mass flow, specific capacity and temperatures of the water-glycol mixture in the dissipation loop. To be congruent with the fact that the thermal capacity is calculated on carbon dioxide side, the evaporator capacity (Eq. (15)) is augmented by the efficiency of the heat exchanger which acts as evaporator in the circuit. As evaporator efficiency was intended the heat actually transferred from the fluid circulating in one side of the heat exchanger to the other. This value allows one to take into account the losses to the environment. For the given heat exchanger, a value equal to 97% was considered as suggested by the manufacturer. In order to compare the parallel compression performance with those of a single-stage cycle operating in similar working conditions the COP of the equivalent single-stage cycle was calculated directly from the data acquired in parallel compression configuration, imagining excluding the parallel compressor and its circuit from the system. It is possible to assume that the cooling capacity of the simpler cycle (Eq. (17)) can be calculated from the difference of the specific enthalpies at the evaporator and gas cooler outlet. It is assumed that the mass flow of carbon dioxide circulating in the circuit is the one that can be provided by the main compressor in the parallel compression cycle.

gcmp1 ¼

Q gc ¼ ðmCO2 ;cmp1 þ mCO2 ;cmp2 Þbhgc;in  hgc;out c

Table 1 Sensors used in the data acquisition system.

The cooling capacity allowed to assess the COP of the equivalent single-stage (ess) cycle (Eq. (18)).

COPess ¼

Fig. 9. Simplified layout of the test rig set-up for parallel compression cycle.

ð17Þ

Q ess W cmp1

ð18Þ

The work that the main compressor would absorb if operated in the equivalent single-stage cycle would be the same measured at the test rig, for it depends only on the gas conditions at its suction and discharge taps. It is important to observe that it is possible to verify the energy balance of the system with a very narrow tolerance once the thermal dissipations along the line between the compressors, high pressure collector and the gas cooler inlet are included. The investigated evaporation pressure spans from 2.25 to 3.25 MPa, the gas cooler outlet temperatures from 22 °C to 40 °C and the discharge pressures from 7 MPa to 11 MPa. The environmental conditions and the design of the dissipation system influenced the range of operating conditions that could be experimented. The evaporation temperature (and pressure) and gas cooler outlet temperature are controlled by means of the dissipation system which uses a PID controller to ensure that the desired temperatures are reached and kept steady during the tests. Because of the large size of the heat exchangers, the temperatures at the gas cooler and evaporator outlets are the same of those controlled by the PID in the dissipation loop. The gas cooler pressure is

A. Chesi et al. / Applied Energy 135 (2014) 274–285

controlled by means of a differential pressure regulator that performs the first part of the lamination of the fluid coming from the gas cooler outlet and a needle pressure regulator, which performs the second part of the lamination of the fluid exiting from the bottom port of the flash tank. In the actual plant the temperature and pressure regulators are not independent and to a modification of one parameter, also other modifications of the others must follow in order to achieve the desired test conditions. Therefore, some conditions were easier to achieve and that is why not for all the evaporation pressures the same amount of tested points can be accounted for. Moreover to estimate the inefficiency of the flash tank as a separator, according to what has been described in Section 3, the separator efficiency was calculated for each test run from the measured data. From the measurements of the mass flow in the section between the vessel and the auxiliary compressor it was possible to infer that from the upper tap of the flash tank exits only vapour. Moreover, the vapour conditions at the compressors inlet are known, being the carbon dioxide temperature and pressure measured directly on the compressor suction tap. Eq. (1) ensures the problem solution when estimating the performance of the theoretical cycle. On the other hand for the measured cycles it is the intermediate pressure measured inside the flash tank which allows the calculation of the separation factor obtained. The error connected to the measurement chain was determined with the analysis method introduced by the authors in [10], as an example, the error on the COP was calculated by means of Eq. (19).

dCOP2 ¼

2  2 @COP @COP @P2ev;out þ @P2gc;out @P ev;out @P gc;out  2  2 @COP @COP þ @T 2ev;out þ @W 2 @T ev;out @W   _ p 2 2 mc þ @T gc;out W

281

Fig. 11. Cooling capacity calculated from measurements at the test rig, the vertical bars indicate the confidence interval of the measurement.



ð19Þ

The average error on the COP of the tested conditions is around 2%.

Fig. 12. Gas cooler capacity calculated by means of data measured at the test rig, the vertical bars indicate the confidence interval of the measurement.

4.2. Results analysis The diagrams in Figs. 10–12 summarize the performance of the parallel compression in flash tank configuration for a series of points whose evaporation mean pressure is 2.65 MPa, the measured points are divided in four different series depending on their gas cooler outlet temperature.

Fig. 10. COP for different discharge pressures and gas cooler outlet temperature as calculated from the data acquired, the vertical bars indicate the confidence interval of the measurement.

The diagram in Fig. 10 shows the trend of the COP as a function of the discharge pressure and gas cooler the outlet temperature. The measured points faithfully duplicate the trend that could be expected from what it is known in literature [8] whereas it cannot be said the same for the absolute values of COP expected mainly because of the critical points covered in this paper. The same considerations can be done for the plots in Figs. 11 and 12 which show the trend of cooling capacity and gas cooler thermal capacity of the same measured points. It is clearly noticeable that not all the conditions tested lead to an improvement in comparison with the equivalent single stage cycle in terms of COP, nevertheless the improvement of cooling capacity is always above 25% but in some cases still far from the theoretical value. The values presented and other measured points which have not been plotted are reported in Tables 2 and 3 and considering some of the parameters provided it is possible to clarify the achieved results. Considering the series of point whose gas cooler outlet temperature is around 40 °C we have three distinct situations. Table 2, presents a series of the most relevant experimental points acquired. For each point are provided test conditions (such as Evaporation pressure, Discharge pressure and Gas cooler outlet temperature) and relevant parameters. (e.g., Separator efficiency, Efficiencies, COP, etc.). The values of intermediate pressure, separator efficiency, cooling capacity and COP of the ideal cycle are calculated by means of the theoretical model being set to account for the

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Table 2 Data measured at the test rig.

a

Point number

Pev, MPa

Pmax, MPa

Pint, MPa

Pint,lima, MPa

Psuc1, MPa

Psuc2, MPa

Tgc,o, °C

Tsuc1, °C

1 2 3 4 5 6 7 8 9 10 11 12

2.46 2.77 2.61 2.60 2.57 2.53 2.61 2.67 2.75 2.74 2.76 2.76

8.06 7.09 8.46 7.78 7.17 9.13 9.34 8.46 7.44 10.91 9.25 8.62

2.94 3.30 3.31 3.12 3.64 3.11 3.01 4.03 3.67 3.06 4.38 6.22

3.45 4.54 4.20 5.24 3.66 4.04 5.25 6.57 4.48 5.11 6.49 5.61

2.39 2.67 2.53 2.52 2.49 2.45 2.53 2.59 2.65 2.66 2.68 2.67

2.64 2.91 2.96 2.74 3.17 2.77 2.71 3.53 3.18 2.78 3.83 4.78

24.9 28.2 30.6 32.4 31.5 31.3 37.3 36.8 36.1 40.1 40.0 39.7

0.3 3.7 0.4 1.4 2.3 3.3 0.3 2.7 1.1 7.5 8.9 9.6

13 14 15 16 17 18 19 20 21 22 23 24

2.17 2.31 2.22 2.22 2.27 2.34 2.32 2.37 2.39 2.30 2.32 2.16

7.95 7.35 9.22 9.44 8.27 8.70 9.59 9.83 10.41 9.18 7.75 8.99

2.54 2.76 2.41 2.41 2.68 2.61 2.51 2.57 2.60 2.58 2.76 2.44

2.97 3.18 3.16 3.15 3.27 3.73 3.45 3.56 3.60 4.11 5.14 4.15

2.11 2.23 2.16 2.16 2.20 2.27 2.25 2.28 2.32 2.24 2.24 2.10

2.30 2.47 2.19 2.19 2.42 2.35 2.28 2.34 2.38 2.33 2.45 2.20

21.9 22.5 24.2 24.3 24.8 28.4 27.4 28.5 29.1 32.6 32.3 32.9

25 26 27 28 29 30 31

3.08 3.29 3.29 3.33 3.20 3.49 3.64

7.07 8.69 9.13 9.50 9.81 7.40 7.16

4.73 4.15 4.05 4.10 3.83 4.53 4.65

4.52 5.69 6.51 6.98 6.41 5.45 4.91

2.96 3.18 3.18 3.22 3.10 3.34 3.46

4.00 3.39 3.42 3.48 3.30 3.91 4.00

30.2 36.8 39.7 41.2 41.6 30.2 28.0

m1, kg s1

m2, kg s1

RV, –

/, –

7.5 6.0 5.0 7.3 3.8 6.4 7.6 0.3 3.6 7.4 3.4 12.3

0.124 0.161 0.137 0.139 0.140 0.123 0.131 0.145 0.154 0.127 0.135 0.138

0.038 0.049 0.045 0.042 0.054 0.040 0.037 0.063 0.055 0.035 0.067 0.099

0.28 0.28 0.28 0.28 0.28 0.28 0.28 0.28 0.28 0.28 0.28 0.28

0.83 0.69 0.76 0.57 1.00 0.74 0.55 0.80 1.00 0.57 0.81 0.81

0.1 1.3 2.9 4.0 0.5 1.6 5.4 5.9 4.5 1.3 3.3 5.9

9.8 8.7 9.7 9.8 9.1 10.0 9.9 9.5 8.7 9.8 9.2 10.2

0.103 0.118 0.100 0.098 0.109 0.113 0.103 0.104 0.106 0.108 0.117 0.103

0.032 0.037 0.027 0.026 0.034 0.031 0.028 0.028 0.028 0.030 0.035 0.028

0.28 0.28 0.28 0.28 0.28 0.28 0.28 0.28 0.28 0.28 0.28 0.28

0.84 0.85 0.74 0.74 0.80 0.67 0.70 0.70 0.70 0.60 0.52 0.56

3.9 13.4 14.0 13.7 13.9 7.2 15.4

5.2 1.1 0.8 0.1 1.7 4.4 5.2

0.194 0.180 0.175 0.178 0.164 0.214 0.216

0.082 0.057 0.057 0.057 0.051 0.077 0.080

0.28 0.28 0.28 0.28 0.28 0.28 0.28

0.96 0.75 0.77 0.87 0.68 1.00 0.94

Tsuc2, °C

Computed (not measured) as described in the paper.

Table 3 Parallel compression cycle performance at the test rig. Point number

W1, kW

W2, kW

g1, –

g2 , –

k1, –

k2, –

Qgc, kW

Qev, kW

Qess, kW

Qev,lim, kW

COP, –

COPess, –

COPlim, –

1 2 3 4 5 6 7 8 9 10 11 12

11.06 11.26 12.44 11.82 11.24 12.77 13.18 12.60 11.66 14.42 13.23 12.69

3.50 3.21 3.73 3.44 3.27 3.90 3.93 3.80 3.38 4.39 4.10 4.04

0.658 0.611 0.627 0.615 0.605 0.625 0.619 0.623 0.615 0.626 0.629 0.630

0.524 0.550 0.529 0.542 0.540 0.526 0.508 0.574 0.550 0.477 0.559 0.552

0.738 0.798 0.753 0.761 0.775 0.721 0.719 0.757 0.788 0.691 0.736 0.756

0.666 0.741 0.677 0.700 0.746 0.653 0.626 0.764 0.749 0.569 0.734 0.866

42.98 43.87 43.24 36.75 21.75 40.43 36.85 33.38 19.35 39.16 36.23 28.44

28.23 30.57 27.68 22.36 6.73 25.20 21.39 18.62 4.73 21.20 19.49 10.23

21.68 23.62 20.98 17.23 5.09 19.12 16.71 13.35 3.72 16.62 13.55 8.03

29.57 33.63 29.82 26.24 6.73 27.32 25.18 19.90 4.58 24.63 20.88 10.72

1.94 2.11 1.71 1.47 0.46 1.51 1.25 1.14 0.31 1.13 1.13 0.61

1.96 2.10 1.69 1.46 0.45 1.50 1.27 1.06 0.32 1.15 1.02 0.63

2.04 2.42 1.87 1.82 0.46 1.65 1.51 1.33 0.32 1.31 1.28 0.63

13 14 15 16 17 18 19 20 21 22 23 24

11.31 11.11 12.19 12.30 11.70 12.20 12.30 12.77 13.15 12.35 11.50 12.00

3.38 3.21 3.61 3.66 3.50 3.59 3.77 3.85 4.01 3.71 3.34 3.87

0.605 0.609 0.612 0.609 0.612 0.608 0.615 0.618 0.623 0.609 0.605 0.610

0.503 0.537 0.458 0.454 0.514 0.484 0.468 0.466 0.465 0.489 0.524 0.475

0.711 0.752 0.683 0.673 0.715 0.706 0.683 0.678 0.670 0.686 0.735 0.688

0.650 0.695 0.587 0.578 0.649 0.608 0.582 0.572 0.562 0.608 0.669 0.611

38.00 40.63 35.85 35.43 38.94 38.16 36.74 36.57 37.13 34.34 32.21 31.23

24.9 27.74 22.94 22.54 25.01 23.69 22.54 22.32 22.70 20.80 18.97 19.40

19.12 21.22 18.16 17.81 19.17 18.70 17.78 17.57 17.95 16.29 14.65 15.27

26.01 28.92 24.63 24.17 26.48 26.13 24.58 24.38 24.75 23.79 22.96 22.61

1.69 1.94 1.45 1.41 1.65 1.50 1.37 1.34 1.32 1.30 1.28 1.22

1.69 1.91 1.49 1.45 1.64 1.53 1.41 1.38 1.37 1.32 1.27 1.27

1.75 2.02 1.52 1.47 1.73 1.65 1.47 1.44 1.42 1.47 1.63 1.43

25 26 27 28 29 30 31

11.31 13.28 13.75 14.25 14.31 11.85 11.50

3.31 3.87 4.01 4.14 4.21 3.44 3.35

0.615 0.647 0.645 0.645 0.648 0.620 0.619

0.531 0.552 0.556 0.555 0.545 0.542 0.525

0.827 0.810 0.792 0.787 0.769 0.853 0.879

0.853 0.719 0.712 0.703 0.677 0.824 0.833

20.08 43.52 39.93 41.01 40.95 29.83 57.66

7.49 24.88 21.40 22.01 22.21 14.65 41.50

6.03 19.23 16.44 16.94 17.12 11.31 30.83

7.55 26.79 22.89 22.57 24.63 14.65 42.17

0.51 1.45 1.21 1.20 1.20 0.96 2.79

0.53 1.45 1.20 1.19 1.20 0.95 2.68

0.51 1.62 1.37 1.32 1.40 1.00 2.89

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actual operating conditions (discharge pressure, evaporation pressure and gas cooler outlet temperature) and the real values of pressure loss and superheating along the lines and compressors efficiency. Once these additional input data are provided, the model solves the cycle and determines the resulting performance for a cycle where the separator efficiency is unitary. Being the separator efficiency and compressors volumetric flow ratio imposed to the model, one of the output variables is the intermediate pressure that should have been measured inside the flash tank if the ideal cycle was achieved on the test rig. In Figs. 13 and 14 the comparisons between theoretical and experimental performance increment in terms of COP and cooling capacity over the ESS cycle are reported, for clarity reasons only the three points described in the discussion hereafter are displayed. Conversely, from what it was done in Tables 2 and 3. the theoretical trends displayed do not take into account the actual pressure losses along the lines because they would differ for each experimental point and the trend would be hard to be interpreted. There are two theoretical lines, dashed and continuous, the first representing the performance expected with the actual separator efficiency achieved in the experimental run, the second is the theoretical limit performance that could be achieved if the separator efficiency had been unitary. The difference between the theoretical actual cooling capacity and the experimental one is usually small and can depend on the absence of pressure losses in the gas cooler when running the theoretical model used. The same difference in terms of COP is influenced also by the different values of the superheating in the compressors inlet lines. Point number 5 in Table 2, at roughly 7.1 MPa of discharge pressure, has a COP which is almost identical to that of the equivalent single stage (0.46 vs 0.45) even if the separator efficiency is 100. In Figs. 13 and 14 the light grey line show the limit COP and cooling capacity achievable by the cycle operated in similar conditions. It can be seen that in the experimental case the performance was close to the one suggested by the theoretical model, the differences are caused by the absence of pressure losses and different superheating levels considered along the lines in the theoretical model when compared to the actual system. In this case, the COP scored is only slightly better than the ESS one because the amount of vapour entering the vessel is very high, therefore it was not possible to reach full liquid conditions out of the bottom tap of the flash tank, consequently the performance obtained is poor. To improve the performance, a higher compressor volumetric flow would be necessary. Point number 8, at about 8.46 MPa gains over 7% on the ESS COP (1.14 vs 1.06) because the amount of vapour entering the flash tank

Fig. 13. COP increase as percentage of the equivalent single stage cycle performance at the test rig for three different points described in the paper. Theoretical limit separation (LS) and Theoretical performance for the actually achieved separation (AS) are shown along with their respective experimental points.

283

Fig. 14. Cooling capacity increase as percentage of the equivalent single stage cycle performance at the test rig for three different points described in the paper. Theoretical limit separation (LS) and Theoretical performance for the actually achieved separation (AS) are shown along with their respective experimental points.

is lower thus the quality at the bottom tap is lower as well, still, if the separator efficiency had been unitary then the COP would have been 1.33 instead of the 1.14 scored. The difference can be observed better in Figs. 13 and 14 where the black dashed line represents the performance expected by the theoretical model when operated in similar conditions to the tested ones. The gap between the filled and empty square points is small. The limit performance increment is far though, as the theoretical model shows with the black continuous line reaching roughly 30% of increment over the ESS COP. Point number 10, at the highest measured discharge pressure of 10.9 MPa, shows a different behaviour: the separator efficiency is low and this causes the cycle not to reach a good COP (1.13) even if the operating condition was an ideal one for the parallel condition and the separator might have been able to actually achieve pure liquid at the bottom port (limit COP was 1.31). This can be observed in Figs. 13 and 14 considering the dark grey dashed and continuous lines and the triangle shaped points, the filled one, that represents the experimental point is below zero in terms of increment of COP over the ESS cycle whereas in terms of cooling capacity matches the actual value calculated by the theoretical model. From Figs. 13 and 14 it is clear that in order to match the theoretical performance investigated in Section 3 and in the literature it is not sufficient to reach a unitary separator efficiency. The pressure loss and heat exchange along the lines of the carbon dioxide circuit are detrimental factors as well, even though less influent when the gas cooler outlet conditions are far from those featured by almost horizontal isotherms on the p–h diagram. From the experimental activity it was possible to carry out an analysis of the parallel compression performance and critical points. The performance reached by the cycle at the test rig was indeed far from the expected one. The main reasons are the differences between the almost ideal cycle which is considered in the thermodynamic analysis and the real cycle that takes place at the test rig. The most important can be summarized as:  First of all the separation, which in the theoretical model is considered complete, in the real system has many factors, such as unsuitable separator design for certain operating conditions and thermal dissipation, that may prevent the separator to function correctly. This causes a big deterioration of the overall performance.  The theoretical model assumes that no pressure loss takes place in any part of the cycle. Not only this pressure losses are unavoidable in the real components but the pressure loss at

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the test bench varies for each single test conditions. In some cases, such as when the discharge pressure is close to the critical value, these losses are very detrimental for performance.  The theoretical model assumes constant ‘‘standard’’ values for the superheating along the suction lines of the compressors, for every operating condition (Pmax, Pev, Tgc,out), and that there is no thermal dissipation towards the environment in the whole system (apart from where it is intended). The real cycle is affected by both highly variable suction lines superheating and thermal dissipations in many of the components of the system, (e.g. the line that travels from the high pressure collector and the gas cooler inlet). This superheating is detrimental for performance because the compressors absorb more power to provide the same compression ratio. During the experimental activity it was noticed that the lower separator efficiency values were registered in those tests where the volumetric flow at the inlet of the flash tank was higher. It is clear that to achieve optimal performance a strong attention must be paid to the separator design.

5. Conclusions The parallel compression cycle in flash tank configuration was analysed by means of a dedicated thermodynamic model. The model allows the analysis of the cycle performance and critical parameters. The study of the expected performance demonstrated the considerable improvements that can be achieved with this cycle layout compared to that with single-stage compression single-throttling. Depending on the operating conditions an ideal cycle may reach COP improvements of over 30% and over 65% in terms of cooling capacity if there is no pressure loss along the line, the superheating is controlled within a certain value (5 K) and the flash tank is considered as a perfect liquid–vapour separator. The analysis emphasizes that the highest improvements are not at the discharge pressure where the maximum COP (or cooling capacity) is reached and proposes a diagram which permits the recognition of the best COP and cooling capacity discharge pressure and its related improvement on the simplest cycle. A further analysis showed the influence of several parameters on the cycle performance. Three main parameters were varied. The first two are the compressors volumetric flow ratio and the intermediate pressure inside the flash tank and they are intimately linked, if the first increase the latter decrease. The third parameter is the separator efficiency that takes into account the inefficiency of the flash tank as separator, this parameter turned out to be very influent on the cycle performance. The final part of the thermodynamic analysis is dedicated to the investigation of the upper and lower limits of the discharge pressure in steady conditions for the cycle proposed. The results showed that these limits vary according to many factors such as the evaporation pressure considered and the compressors volumetric flow ratio. Lower evaporation pressures widen both upper and lower limits, whereas lower compressors volumetric flow ratios tend to narrow the lower limit and widen the higher one. The experimental activity which followed the thermodynamic analysis was carried out on the test rig dedicated to carbon dioxide refrigeration cycles and components investigation of the Department of Industrial Engineering of the University of Florence. The tested conditions span different evaporation pressure, discharge pressure and gas cooler outlet temperature ranges. Most of the tests are at typical conditions for real cooling applications. Significant performance parameters were determined: compressor volumetric efficiency and energy efficiency, cycle cooling capacity and COP of the parallel and equivalent single stage cycle, thermal

capacity and actual separator efficiency. The parallel compression cycle performance plots confirm the expected trend for each parameter but the absolute value is lower. From the data acquired it was possible to estimate the detrimental effects on performance of several factors. The analysis confirms what was observed in the theoretical analysis about the separator efficiency, it is crucial that the liquid and vapour are effectively separated inside the flash tank, indeed a reduction of the separator efficiency is particularly detrimental in the conditions where the flash tank inlet has a high quality (high percentage of gas). Another detrimental factor is represented by the pressure losses along the lines of the system. Suction line pressure loss contributes to increase the work absorbed by compressors, anyway the losses in the gas cooler inlet line and inside the gas cooler itself may lead to strong performance reduction if the gas cooler outlet temperature and discharge pressure are close to the values where the isotherms are almost horizontal on the p–h diagram, just above the critical point. The last parameter which concurs to undesirable performance losses is the thermal superheating, especially along the suction lines of both compressors where they lead to an increase of compressor power absorption because of the trend of the isentropic transformation lines which tend to be less sloping as the temperature rises. Considering all of these effects, the results obtained suggest that the most relevant factors for the components design are low pressure losses along the piping, a good insulation (especially of the low temperature side of the circuit) and an optimal design of the separator. Moreover, in order to achieve the maximum theoretical performance increase for this particular layout in different conditions it is highly recommended to use compressors whose volumetric flow ratio can be modified. References [1] Lorentzen G. Revival of carbon dioxide as a refrigerant. Int J Refrig 1994;17:292–301. [2] Cavallini A, Nekså P. Prospects for the return of CO2 as a refrigerant. Buenos Aires (Argentina): CIAR; 2001. p. 761–90. [3] Cecchinato L, Corradi M, Fornasieri E, Minetto S, Zilio C, Schiavon, A. Theoretical and experimental analysis of a CO2 refrigerating cycle with twostage throttling and suction of the flash vapour by an auxiliary compressor. In: IIR International conference on thermophysical properties and transfer processes of refrigerants. Vicenza (Italy); 2005. [4] Chesi A, Ferrara G, Ferrari L, Tarani F. Setup and characterisation of a multipurpose test rig for R744 refrigerating cycles and equipment. Int J Refrig 2012;35:1848–95. [5] Kim MH, Pettersen J, Bullard CW. Fundamental process and system design issues in CO2 vapor compression systems. Prog Energy Combust Sci 2004;30:119–74. [6] Cabello R, Sanchez D, Llopis R, Torrella E. Experimental evaluation of the energy efficiency of a CO2 refrigerating plant working in transcritical conditions. Appl Therm Eng 2008;28:1596–604. [7] Cavallini A, Cecchinato L, Corradi M, Fornasieri E, Zilio C. Two-stage transcritical carbon dioxide cycle optimisation: a theoretical and experimental analysis. Int J Refrig 2005;28:1274–83. [8] Cho H, Ryu C, Kim Y. Cooling performance of a variable speed CO2 cycle with an electronic expansion valve and internal heat exchanger. Int J Refrig 2007;30:664–71. [9] Aprea C, Maiorino A. An experimental evaluation of the transcritical CO2 refrigerator performance using an internal heat exchanger. Int J Refrig 2008;31:1006–11. [10] Rigola J, Ablanque N, Pérez-Segarra CD, Oliva A. Numerical simulation and experimental validation of internal heat exchanger influence on CO2 transcritical cycle performance. Int J Refrig 2010;33:664–74. [11] Torrella E, Sánchez D, Llopis R, Cabello R. Energetic evaluation of an internal heat exchanger in a CO2 transcritical refrigeration plant using experimental data. Int J Refrig 2011;34:40–9. [12] Agrawal N, Bhattacharyya S. Studies on a two stage transcritical carbon dioxide heat pump cycle with flash intercooling. Appl Therm Eng 2007;27:299–305. [13] Bell, I. Performance increase of carbon dioxide refrigeration cycle with the addition of parallel compression economization. In: 6th IIR Gustav–Lorentzen conference on natural working fluids. 2/A4.30, Glasgow (UK); 2004. [14] Agrawal N, Bhattacharyya S, Sarkar J. Optimization of two stage transcritical carbon dioxide heat pump cycles. Int J Refrig 2007;46:180–7. [15] Sarkar J, Agrawal N. Performance optimization of transcritical CO2 cycle with parallel compression economization. Int J Therm Sci 2010;49:838–43.

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