Second-Law Analysis to Improve the Energy Efficiency of Screw Liquid Chillers

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1 Entropy 2010, 12, ; doi: /e OPEN ACCESS entropy ISSN Article Second-Law Analysis to Improve the Energy Eiciency o Screw Liquid Chillers zong-shing Lee Department o Energy and Rerigerating Air-Conditioning Engineering, National aipei University o echnology, aipei 10608, aiwan; tslee@ntut.edu.tw Received: 1 January 2010; in revised orm: 2 February 2010 / Accepted: 1 March 2010 / Published: 4 March 2010 Abstract: his work applies the second-law analysis o thermodynamics to quantiy the exergy destruction o the components o screw liquid chiller, and to identiy the potential or each component to contribute to improve the overall energy eiciency o the system. hree screw liquid chiller units were built to demonstrate the easibility o the model presented herein. Unit A was a 100 R water-cooled screw liquid chiller. Unit B was modiied rom Unit A by switching the old condenser or a new one with a greater heat transer, and Unit C was modiied rom Unit B by exchanging the compressor or a more eicient one. he results indicate that the compressor has the largest potential to improve energy eiciency, ollowed in order by the condenser, and then the evaporator. he second law analysis may help engineers to ocus on the components with higher exergy destruction and quantiy the extent to which modiying such components can inluence, avorably or unavorably, the perormance o other components o the screw liquid chiller. Keywords: second-law analysis; screw liquid chiller; energy eiciency; irreversibility; exergy destruction Nomenclature c p speciic heat (kj/kg K) COP coeicient o perormance g gravitational acceleration (m/s 2 ) h enthalpy (kj/kg) I exergy destruction or irreversibility (kw)

2 Entropy 2010, m mass low rate (kg/s) Q heat transer rate (kw) R rerigerating tons s speciic entropy (kj/kg K) S entropy generation (kj/k) gen temperature V velocity W work z elevation Greek symbols η II second-law eiciency Subscripts 1,2,7 components act actual ci, co cooling water inlet and outlet, respectively comp compressor cond dis exp evap H i in L liq condenser discharge expansion valve evaporator rerigerant high independent variable (i=1,2,,n) inlet low liquid o surrounding out out rev reversible suc suction wi, wo chiller water inlet and outlet, respectively 1. Introduction Vapor-compression liquid chillers have been commonly used to cool water, brine and other secondary coolants, in commercial and industrial air-conditioning or rerigeration systems. he primary components o a vapor-compression liquid chiller include a compressor and its driver, a condenser, a throttling device, a liquid cooler (evaporator), and a control system. Speciications o the energy eiciency o liquid chillers typically involve the coeicient o perormance (COP), the energy eiciency ratio (EER) and the input energy ratio (kw/r) as indices. All such approaches to evaluating the energy eiciency o liquid chillers are based on the irst law o thermodynamics. he measured quantity is thus called irst-law eiciency. However, the irst law o thermodynamics states

3 Entropy 2010, the conservation o energy and the transormation o energy rom one orm to another. Preserving the quality o energy and increasing the energy eiciency o liquid chillers are major concerns to engineers, and the second law provides the necessary means to determine the quality as well as the extent o degradation o energy during a process. he energy eiciency o a liquid chiller is a consequence o the destruction o available energy (or exergy) contributed by every system component, so the second-law analysis, incorporating the irst and second laws o thermodynamics, directly evaluates the potential to improve the eiciency o liquid chillers. Recent studies have increasingly applied the second-law analysis in the ields o rerigeration, air-conditioning and heat pump systems [1 13]. Second-law analysis can be divided into exergy (or availability) and irreversibility analyses. he ormer addresses the conversion and loss o exergy while the latter concern the entropy generation and irreversibility. Bejan [14] detailed undamental theories and concepts. Liang and Kuehn [1] perormed irreversibility analysis on data obtained rom a R-22 reciprocating water-cooled liquid chiller with a cooling capacity o 7.56 kw in a steady state. he analytical results show that the percentages o the irreversibility associated with the components were, compressor 44.41%, condenser 27.11%, evaporator 18.13%, expansion valve 4.06%, suction line 4.44% and discharge line 1.85%. ASHRAE [8] illustrated the second-law analysis o an R-22 reciprocating, air-cooled, direct expansion rerigerator, with a cooling capacity o 7.0 kw. he results indicated that the percentages o the total irreversibility associated with the components were compressor 46%, condenser 18%, expansion valve 18%, evaporator 9%, discharge line 5%, suction line 3% and liquid line 0%. he results dier rom those o Liang and Kuehn [1]. Brian et al. [9] applied the second law analysis to quantiy the irreversibility o components in a household reezer with a volume o 18 t 3 and an air-conditioner with a capacity o 10.5 kw. hey identiied the potential o every component to improve the overall energy eiciency o the system. Speciically, their analysis revealed that the percentage irreversibilities associated with the components o the reezer ollowed the order o the compressor, the condenser and then the evaporator. his sequence diers rom that obtained or the air-conditioner, or which the evaporator ranked irst, ollowed by the condenser and the compressor. Yumrutaş et al. [11] used exergy analysis to develop a theoretical model to investigate the eects o the evaporating and condensing temperatures on the pressure drop, the exergy destruction, the second-law eiciency and the coeicient o perormance (COP) o a vapor compression rerigeration cycle. he results revealed that the two design parameters o evaporating and condensing temperature strongly aected the exergy destructions associated with the evaporator and the condenser, the second-law eiciency and the COP o the cycle, but only weakly inluenced the exergy destructions in the compressor and the expansion valve. Reducing the dierence between the temperature o the evaporator and that o the rerigerated space, and between the temperature o the condenser and that o the outside air can reduce the total loss o exergy and simultaneously improved the COP and second-law eiciency. he works cited above shows that the application o the second-law analysis to commercial screw liquid chillers is still lacking. he main goal o this study is to develop a methodology, based on second-law analysis, to be applied to evaluate the irreversibility o the components o a screw liquid chiller, and to identiy the potential or each component to contribute to the energy eiciency o the overall system. hree screw liquid chillers, units A, B and C, are built to conirm the model developed

4 Entropy 2010, herein. Unit A is selected as a baseline unit in this work. It is a liquid chiller with a nominal cooling capacity o 100 tons and screw type compressor; it has a direct-expansion plate-type evaporator and a water-cooled, shell-and-tube type condenser. Unit B is the chiller modiied rom unit A by replacing the old condenser with a new one with enhanced heat-transerring cooper tubes, and Unit C is modiied rom Unit B by replacing the compressor with a more eicient one. Comparisons were made to evaluate the improvement by second-law analysis. Engineers could use the analytical method presented in this study to quantiy the irreversibility o components in their liquid chiller designs and provide a clear direction in which to increase the system energy eiciency o a liquid chiller. he analysis also provides insights into how eicient components inluence the potential payo o improving another component. 2. Second-Law Analysis 2.1. hermodynamic Modeling he actual vapor-compression rerigeration cycle is the most extensively employed cycle in liquid chillers. his actual cycle consists o several components, such as a compressor, a condenser, an expansion device, an evaporator and connecting pipelines. It diers rom the ideal vapor-compression rerigeration cycle in many ways, mostly due to the irreversibility associated with various components. wo common sources o irreversibility in the actual rerigeration cycle are luid riction, which causes alls in pressure, and heat transer across a inite temperature dierence. Consider each component in a real vapor-compression cycle, as indicated in Figure 1, as a control volume that undergoes a steady-low process. he irst and second laws o thermodynamics or a steady-low process are as ollows [14]: n i= V V W + m ( h + + gz) m i ( h + + gz) = 0 (1) 2 2 in n i= 0 i i out + ms ms + S in he irreversibility (or exergy destruction) o each process can be determined using the Guoy- Stodola relation: out gen = 0 I = S o gen (3) he energy and entropy balance can be simpliied by making urther assumptions or each component. he evaporator, condenser, expansion valve and connecting pipelines are assumed to do no work. he throttling process o the expansion device is assumed to be adiabatic. he compressor perorms a real compression process. he changes in kinetic and potential energy o all components are negligible. (2)

5 Entropy 2010, Figure 1. Schematic diagram o vapor-compression liquid chiller. ci co H Q cond liq Q dis comp W comp suc Q evap L wo wi 2.2. Second-Law Analysis o Chiller Components he control volume o the evaporator, chosen or process analysis includes the entire evaporator, except that occupied by chilled water. In this process, heat is transerred rom the chilled water to the rerigerant. he heat transer rate Q evap, the rerigerant mass rate m, and the rate o entropy generation, can be determined as ollows: S gen, evap S Q ( ) evap wc p wi wo (4) gen, evap m evap = ( h 1 h7) ( s1 s7) In Equation (6), L is the mean temperature during the transer o heat in the evaporator, and is determined by the ollowing relation: L wi = wi ln where wi and wo are the temperatures o the chilled water that enters and leaves the evaporator, respectively. wo wo evap L (5) (6) (7)

6 Entropy 2010, he compression process is assumed to be non-adiabatic and irreversible. In this process, the energy across the control surace including a heat loss rom the compressor to its surrounding, and a work input rom the surrounding to the compressor. he rates o heat loss and entropy generation in the compression process can be determined as ollows: comp ( h3 h2) + W comp (8) Qloss S gen, comp ( s3 s2) where o is the temperature o the surroundings, and W comp is the work input to the compressor, which can be experimentally measured. he control volume o the condenser used in the analysis includes the entire volume o the condenser except that occupied by cooling water. Heat is transerred rom the rerigerant to the cooling water. he heat transer rate, Q cond, rom the rerigerant to the cooling water, and the rate o entropy generation can be determined using: S gen, cond S gen, cond o Q ( h ) cond 5 h4 (10) ( s5 s4) where H is the mean temperature o cooling water, which can also be treated as a mean temperature in the heat transer process. It can be evaluated as ollows: H co = co ln where ci and co represent the inlet and outlet temperatures, respectively, o the cooling medium in the condenser. he process involving in the throttling device is assumed to be an isenthalpic process. hereore, the ollowing relation holds: ci ci cond H (9) (11) (12) h 7 = h 6 (13) ( s ) 6 S gen, exp 7 s In this work, the pipelines considered include the suction, the discharge and the liquid lines, but not the line between the throttling device and the evaporator, since the throttling device and the evaporator are normally located very close to each other. he heat gained or heat lost between the pipelines and their surrounding, and the associated rate o entropy generation can be expressed as ollows: S (14) Q ( h ) 2 h1 (15) gen, suc suc ( s2 s1) suc o (16)

7 Entropy 2010, S S gen, dis gen, liq dis ( h ) 4 h3 ( s4 s3) liq ( h ) 6 h5 ( s6 s5) dis o liq o (17) (18) (19) (20) hus, the total exergy destruction or irreversibility o a chiller is: I = o ( Sgen, evap + Sgen, suc + Sgen, comp + Sgen, dis + Sgen, cond + Sgen, liq + Sgen,exp) i (21) he exergy destruction ratio or each component o liquid chiller can be deined as ollows: y d I i = I i (22) 2.3. First- and Second-Law Eiciency o Chillers he coeicient o perormance o a vapor-compression chiller unit, deined based on only the irst law o thermodynamics, is sometimes reerred to as the irst-law eiciency, and is deined as ollows: COP act evap = (23) W comp he second-law eiciency η II, can be deined as the ratio o the actual COP to the maximum possible COP under the ixed conditions: where COPrev is deined as: COP η act II = (24) COPrev COP L rev = (25) ( H L ) In Equation (25), L and H are the absolute low- and high- temperature limits, respectively, as deined in Eqs. (7) and (12). 3. Experimental System and Method 3.1. Chiller Units and Experimental System hree sets o chiller units, abbreviated as Units A, B and C, were built and tested under the same conditions to validate the eectiveness o the second-law analysis. Unit A is a typically designed R-22 water-cooled screw liquid chiller with a nominal capacity o 100 R. Figure 2 schematically depicts the system layout. he compressor is a semi-hermetic twin-screw compressor with a displacement o

8 Entropy 2010, m 3 /h. he evaporator is a direct-expansion plate-type heat exchanger. he water-cooled condenser is o the shell-and tube type and made o copper tubes whose exterior suraces are inned and whose interior suraces are smooth. he throttling device was a thermostatic expansion valve with an external equalizer. Figure 2. he measured points o screw liquid chiller. A: Evaporator ci : Entering cooling water temperature o condenser B: Suction line co : Leaving cooling water temperature o condenser C: Screw compressor wi : Entering chilled water temperature o evaporator D: Discharge line wo : Leaving chilled water temperature o evaporator E: Condenser o : Surrounding temperature F: Liquid line F c : Condenser cooling water low rate G: Expansion valve F w : Chilled water low rate 1-7 : Rerigerant-side temperatures Q: Heat transer P 1 -P 7 : Rerigerant-side pressures W: Work Unit B was the same as Unit A, except in that its condenser had been modiied by replacing the externally inned and internally smooth copper tubes with externally inned and internally spiral copper tubes, to enhance the heat transer on the water-side. Finally, Unit C was the same as Unit B, except in that Unit B s compressor was replaced by a more eicient one with the same displacement. he subcooling temperature o the condenser and the superheating temperature o the evaporator were set to 5 K or all chiller units.

9 Entropy 2010, he system used to conduct the chiller tests can supply both chilled- and cooling-water at a low rate steady to within ± 1%, and at a temperature maintained within ± 0.1%. Four physical parameters were measured - temperature, low rate, pressure and power input to the compressor. All the measurement instruments conormed to the requirements o ARI 550/590 [15]. Figure 2 presents the position o the mounting o each measuring instrument, while able 1 states the accuracy o the respective instruments. able 1. Measurement instruments and accuracy. Physical parameter Instrument Accuracy emperature Platinum resistance thermometer ± 0.03 C Pressure Pressure transducer ± 0.15% Flow rate Electromagnetic low meter ± 0.5% Power consumption Wattmeter ± 0.5% 3.2. Experimental Methods he testing conditions included setting the temperature o the chilled water that entered and let the unit to 12 ± 0.5 C and 7 ± 0.5 C, and setting the temperature o the cooling water that entered and let the unit to 30 ± 0.5 C and 35 ± 0.5 C. Once the liquid chiller unit reached a steady state and had been kept running or over one hour, physical measurements were taken at 20 minute intervals, our times. he averages o these our sets o data were used in the second-law analysis to quantiy irreversibility. In each test, the heat balance was within 5% per ARI Standard 550/590 [15]. Each test was careully perormed at the same temperature o the surroundings in the laboratory, 25 C, to increase the reliability o the test results. 4. Results and Discussion able 2 presents the measured temperature and pressure data at each state point in the three chiller units under the same testing conditions, rom which the enthalpy, entropy and other thermodynamic properties at each state could be determined. Figure 3 plots a pressure-enthalpy diagram o the vapor compression rerigeration cycle o each chiller unit. he second-law or irreversibility analysis was conducted using the data in able 2 and the related equations introduced in the theoretical section. able 3 presents the results o the second-law analysis. Figure 4 compares the exergy destruction o each component or the three chiller units studied in the work. Figure 5 plots the exergy destruction ratio o components or three units. he ollowing discussion o the analytical results is based on able 3 and Figures 3 to Perormance o the Baseline Chiller (Unit A) As shown in able 3, the irst-law analysis or Unit A reveals a COP o 3.7. Additionally, the compressor holds the greatest entropy generation, ollowed in order by the condenser and the evaporator, with the entropy generation o , and kw/k, respectively. Figure 4 shows the comparison o exergy destruction associated with components or Units A, B, and C. In the

10 Entropy 2010, case o Unit A, the compressor exhibited the largest exergy destruction with kw, ollowed by the condenser with kw and the evaporator with kw. Figure 5 depicts that the exergy destruction ratios were 45%, 27% and 17% or the compressor, the condenser and the evaporator, respectively. he sum o the exergy destruction ratios o these three components approaches 90%. Apparently, these three components should be considered irst in increasing the energy eiciency o the screw liquid chiller. he second-law eiciency η II o Unit A is around 30.27%. Figure 3. Pressure-enthalpy diagram o chiller units under the same testing conditions. 2 P [MPa] 1 Unit A Unit B Unit C h [kj/kg] Figure 4. Comparison o exergy destruction associated with components or Units A, B and C. 70 Exergy Destruction [kw] Expansion valve Liquid line Condenser Discharge line Compressor Suction line Evaporator 0 Unit A Unit B Unit C

11 Entropy 2010, Figure 5. Comparison o exergy destruction ratio associated with components or Units A, B and C. 100% 90% Expansion valve Exergy Destruction Ratio [%] 80% 70% 60% 50% 40% 30% 20% Liquid line Condenser Discharge line Compressor Suction line 10% 0% Unit A Unit B Unit C Evaporator 0.00 able 2. hermodynamics properties measured and calculated at various state points. Chiller Unit State Point [ ] P [kpa] sat [ ] h [kj/kg] s [kj/kgk] 1. Evaporator outlet Compressor inlet Compressor outlet A 4. Condenser inlet Condenser outlet Expansion valve inlet Evaporator inlet Evaporator outlet Compressor inlet Compressor outlet B 4. Condenser inlet Condenser outlet Expansion valve inlet Evaporator inlet Evaporator outlet Compressor inlet Compressor outlet C 4. Condenser inlet Condenser outlet Expansion valve inlet Evaporator inlet

12 Entropy 2010, able 3. Results o irst-law and second-law analysis or three screw liquid chiller units. Unit Component First-law analysis Second-law analysis Heat transer Work First law eiciency Entropy generation Exergy destruction Q [kw] W [kw] COP S gen [kw K -1 ] I [kw] Exergy destruction ratio y d [%] Second law eiciency η II A B C Evaporator % Suction line % Compressor % Discharge line % 3.7 Condenser % Liquid line % Expansion valve % otal % Evaporator % Suction line % Compressor % Discharge line % 4.1 Condenser % Liquid line % Expansion valve % otal % Evaporator % Suction line % Compressor % Discharge line % 4.5 Condenser % Liquid line % Expansion valve % otal % Eect o Improving Condenser Heat ranser on Cop (Unit B) he analytical results in able 3 show that the COP o the rerigeration cycle is 4.1, which is approximately 10.8% higher than that o Unit A. he entropy generation or each component ollows the order o compressor, condenser and evaporator, with values o 0.084, and kw/k, respectively, ollowing the same order as or Unit A. Examining Figure 4 and 5, we ind that the highest exergy destruction is that o the compressor at kw, ollowed by that o the condenser at kw and then that o the evaporator at kw. he exergy destruction ratios or individual component are 47%, 22% and 19%, respectively. Moreover, the sum o the exergy destruction o these three components approaches 88% o the whole unit. he second-law eiciency η II o Unit B is 33.2%, representing an improvement o around 9.7% over Unit A. Closely examining Figure 3 and able 3 yield urther important results. Figure 3 compares the rerigeration cycles o the three chiller units and indicates that Unit B exhibited a lower condensing temperature and discharge temperature than that o Unit A. Since Unit B has the same construction as Unit A, except in that its condenser had been modiied by replacing the externally inned and internally smooth cooper tubes with externally inned and internally spiral cooper tubes, to enhance the

13 Entropy 2010, heat transer on the internal water-side. When the condenser in Unit A was replaced with a new condenser, the heat transer eect on the water side was increased by the structure o the externally inned and internally spiraling copper tube, reducing the dierence between the condensing temperature and that o cooling water, and thus reducing the power consumption o the compressor. hereore, the entropy generation and the exergy destruction associated with the condenser were reduced, reducing the total exergy destruction and improving the energy eiciency o the Unit. Synthetically analyzing and comparing data concerning each item indicates that replacing the condenser practically promotes heat transer and increases the energy eiciency o the unit Eect o Improving Compressor Eiciency on Cop (Unit C) he analytical results or Unit C shown in able 3 indicate that the COP o the rerigeration cycle is 4.5, which are about 9.8% higher than o Unit B and 20.4% higher than that o Unit A. he amount o entropy generation o each component ollows the same order as that or Units A and Unit B- the compressor, ollowed by the condenser, then by the evaporator, with respective values o , and As displayed in Figure 4 and Figure 5, the values o exergy destruction associated with each component in this regard ollow the same order as that or Units A and B. he component with the highest exergy destruction remains the compressor at 19.2 kw, ollowed by the condenser at kw and the evaporator at kw, yielding exergy destruction ratio o 38%, 25% and 23%, respectively. he summation o these three values is 86% o the total irreversibility, which is lower than that or Unit B. However, the second-law eiciency is urther increased to 36.5%, representing improvements o around 20.5% and 10% over that o Unit A and that o Unit B, respectively. As comparing the rerigeration cycles o the chiller units indicates that the discharge temperature o Unit C was lower than that o Unit B, and the compression process or Unit C was near to isentropic process. Since Unit C was constructed by replacing the compressor in Unit B with a more eicient one, so the compression process o Unit C is more nearly to become isentropic than that o the other two Units, signiicantly reducing the generation o entropy and the irreversibility o the compressor, thus reducing the total irreversibility. he energy eiciency o the unit is thus increased. All these data reveals that replacing a compressor with one o higher compression eiciency is easible, and it is particularly helpul in raising the energy eiciency o the liquid chiller. Second-law analysis explicates the exergy destruction associated with each component in a liquid chiller. It clearly indicates which component is most able to increase the energy eiciency. he analytical method presented herein can be applied not only to evaluate the eectiveness o modiications o components, but also to determine how the eiciency o a component inluences the potential payo o improving another component. Hence, this work has practical value and the method developed in this study can be used as an R&D tool or improving the energy eiciency o screw liquid chillers. 5. Conclusions his work established a method or identiying the potential o each component in a screw liquid chiller to improve the energy eiciency o chiller by applying second-law analysis. hree liquid chillers were constructed or testing and analysis. he improvements o eiciency and the

14 Entropy 2010, appropriateness o second-law analysis were clariied in a case study. he results support the ollowing conclusions. he exergy destruction or irreversibility associated with each component in a screw liquid chiller taken rom the testing unit ollow the sequence, compressor associated with 38% to 47% o the total system irreversibility, ollowed by the condenser associated with 22% to 27% thereo, ollowed by the evaporator associated with 17% to 23%. A component associated with a higher proportion o exergy destruction may more dramatically contribute to improve energy eiciency o a screw liquid chiller. Reducing the exergy destruction associated with components may increase the system COP and the second-law eiciency. Second-law analysis may help us to identiy components with higher exergy destruction, modiications to which have greater potential to increase eiciency. Second-law analysis is demonstrated to be a easible and eective means in improving the energy eiciency o a screw liquid chiller. Acknowledgements he authors would like to thank the National Science Council o aiwan or inancially supporting this research under Contract No. NSC E CC3. Reerences 1. Liang, H; Kuehn.H. Irreversibility analysis o a water to water mechanical compression heat pump. Energy 1991, 16, McGovern, J.A.; Harte, S. An exergy method or compressor perormance analysis. Int. J. Re. 1995, 181, Zubair, S.M.; Yaqub, M.; Khan, S.H. Second-law-based thermodynamic analysis o two-stage and mechanical-subcooling rerigeration cycles. Int. J. Re. 1996, 19, Domanski, R.; Fellah, G. hermoeconomic analysis o sensible heat, thermal energy storage systems. Appl. herm. Eng. 1998, 18, Moran, M.J. On second-law analysis and the ailed promise o inite-time thermodynamics. Energy 1998, 23, Lee, S.F.; Sheri, S.A. Second law analysis o a lithium bromide/water absorption system or cooling and heating applications. Presented at the 20th International Congress o Rerigeration, Sydney, Australia, 19th 24th September, Meunier, F. Second-law analysis o rerigeration cyclic open systems. ASHRAE rans. 1999, 105, ASHRAE Fundamentals Handbook; ASHRAE: Atlanta, GA, USA, Bridges, B.D.; Harshbarger, D.S.; Bullard, C.W. Second law analysis o rerigeration and air conditioners. ASHRAE ransactions 2001, 107, Bilgen, E; akahashi, H. Exergy analysis and experimental study o heat pump systems. Exergy 2002, 2, Yumrutaş, R; Kunduz, M; Kanoğlu, M. Exergy analysis o vapor compression rerigeration systems. Exergy 2002, 2,

15 Entropy 2010, Ozgener, O.; Hepbasli, A. Experimental perormance analysis o a solar assisted ground-source heat pump greenhouse heating system. Energ. Bldg. 2005, 37, Esen, H.; Inalli, M.; Esen, M.; Pihtili, K. Energy and exergy analysis o a ground-coupled heat pump system with two horizontal ground heat exchangers. Bldg. Environ. 2007, 42, Bejan, A. Advanced Engineering hermodynamics; John Wiley & Sons: New York, NY, USA, ARI. Water Chilling Packages Using the Vapor Compression Cycle; ARI: Arlington, VA, USA, by the authors; licensee Molecular Diversity Preservation International, Basel, Switzerland. his article is an open-access article distributed under the terms and conditions o the Creative Commons Attribution license (