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  • 1 Department of Energy Engineering, Faculty of Mechanical Engineering, Budapest University of Technology and Economics, Műegyetem rkp. 3, H-1111Budapest, Hungary
  • | 2 Technical College of Kirkuk, Northern Technical University, 36001Kirkuk, Iraq
  • | 3 Department of Thermohydraulic, Centre for Energy Research, POB. 49, H-1525Budapest, Hungary
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Abstract

Thermodynamic efficiency is a crucial factor of a power cycle. Most of the studies indicated that efficiency increases with increasing heat source temperature, regardless of heat source type. Although this assumption generally is right, when the heat source temperature is close to the critical temperature, increasing the heat source temperature can decrease efficiency. Therefore, in some cases, the increase in the source temperature, like using improved or more collectors for a solar heat source can have a double negative effect by decreasing efficiency while increasing the installation costs. In this paper, a comparison of the CO2 subcritical cycle and the Trilateral Flash Cycle will be presented to show the potential negative effect of heat source temperature increase.

Abstract

Thermodynamic efficiency is a crucial factor of a power cycle. Most of the studies indicated that efficiency increases with increasing heat source temperature, regardless of heat source type. Although this assumption generally is right, when the heat source temperature is close to the critical temperature, increasing the heat source temperature can decrease efficiency. Therefore, in some cases, the increase in the source temperature, like using improved or more collectors for a solar heat source can have a double negative effect by decreasing efficiency while increasing the installation costs. In this paper, a comparison of the CO2 subcritical cycle and the Trilateral Flash Cycle will be presented to show the potential negative effect of heat source temperature increase.

1 Introduction

The demand for energy in the world increases; this increase is preferably satisfied by modern power plants using a clean, renewable source. A special class of these power plants is the one operated by low heat sources, using organic Rankine cycles or carbon dioxide power cycle. The transcritical CO2 power cycle (working with heat sink temperature below and heat source temperature above the critical temperature (Tcr) of CO2, namely 31 °C) has significant and sufficient potential to convert the heat to produce power (electricity) by using the carbon dioxide as a working fluid due to its good thermodynamic and environmental properties [1, 2]. Many thermodynamics cycles are applicable in the temperature below 350 °C, like the CO2 transcritical power cycle, Organic Rankine Cycle (ORC) and Trilateral Flash Cycle (TFC), instead of high-temperature steam Rankine Cycle (RC) [1, 3–5]. CO2 has a low critical temperature, which makes it appropriate for utilizing a low heat source – geothermal [6, 7]. CO2 power cycles are widely used in air conditioning, heat pumps, refrigerating systems, and power cycles [8–11]. CO2 power cycle can be used to utilize solar heat, but due to the weather-dependence of this source, a heat storage system is required to provide continuous heating. Also, integration of the absorption refrigeration system with reheat transcritical power cycle leads to improve the efficiency and maintain stable productivity by keeping the low condensation temperature at all different weather conditions. Using the compressed air energy storage help to overcome the decrease or interruption for solar energy and improves the technical flexibility in solar thermal power and storage [12, 13]. Using a regenerative heat exchanger (recuperator) in the CO2 transcritical cycle improves overall net power and overall thermal efficiency [14]. At the low and high heat source temperature, the transcritical CO2 power cycle using mixed CO2 is better than the cycle that using pure carbon dioxide, thermodynamically, and exergo-economically. It has been shown by analyzing binary mixtures of carbon dioxide with other refrigerants (R32, R1270, R161, R1234yf, R1234ze, R152a), and alkanes (butane, pentane, propane, isobutene or isopentane) that the highest exergy efficiency and the lower levelized cost per unit of exergy product was with CO2/propane at the high-temperature heat source. At lower heat source temperature, the highest exergy efficiency was with CO2/R32, and the lowest levelized cost per unit of exergy product was with CO2/R161 [15]. With CO2 mixtures that consist of binary mixture of CO2 with one of these refrigerants (R152a, R161, R290, R1234yf, R1234ze, and R1270) with a transcritical power cycle at geothermal water temperature between 100 and 150 °C and temperature of cooling tower 10–30 °C, it has been observed that the better thermal performance and economic performance was with R161/CO2, while the R290/CO2 was the worst due to low thermal performance. The cost per net power reduction, decrease of operating pressure and extension of the range of condensing temperature, all these occur with the blends of CO2 more than the CO2 in a pure state. At the low cooling water temperature, R152a/CO2 mixed working fluid is not suitable with the proposed system [16]. The comparison between basic, recuperator, reheat, and regenerative systems for transcritical CO2 power cycle demonstrated that reheat transcritical CO2 cycle is the best one, concerning thermo-economic performance, and reheat system showed an increase in net power produced, energy efficiency, and efficiency of exergy compared the basic transcritical CO2 cycle while the total investment cost is higher for reheat system due to largest heat transfer area [17]. Condensing is one of the problems facing the conventional CO2 trans- and subcritical CO2 power cycles by using traditional water cooling, but with self-condensing, the CO2 carbon dioxide transcritical power cycle overcome this problem and can operate well with the cooling water as warm as 30 °C [18, 19]. The CO2 transcritical power cycle has a better economic performance than the organic Rankine cycle in terms of cost per net power output and under a certain turbine inlet pressure. The cost per net power output in the regenerative CO2 transcritical power cycle is even lower than that of the basic CO2 transcritical power cycle, that which observed by analyzing the organic Rankine cycle and CO2 transcritical cycle with a geothermal heat source and different working fluids for example isobutane, R123, pentane R245fa [20].

The subcritical CO2 power cycle (CO2 Rankine cycle) can use the ambient temperature or low-temperature geothermal as a heat source; these two sources are classified as a low enthalpy source [21]. To properly characterize the transcritical cycle, it requires a deep knowledge of the subcritical carbon cycle [22]. When the ambient temperature or some other source with similar temperature (for example, thermal water not above 35–40 °C) are used as a heat source, then - depending on the weather conditions - transcritical cycles might shift to the sub-critical region (i.e., the maximum temperature will be below the critical temperature of CO2). In that case, it is better to use a subcritical CO2 cycle. Still, engineers should know that there is a narrow temperature range near to the critical temperature, where the thermodynamic efficiency has inverse maximum cycle temperature dependence. Therefore, there is a temperature range, which should be avoided during this application. The aim of the study is to demonstrate to engineers and researchers that the efficiency does not always increase with increasing heat source temperature, but sometimes the increasing of the heat source temperature close to the critical temperature leads to decreasing the efficiency. That can also happen with other thermodynamics cycles using various working fluids, for example Rankine cycle and organic Rankine cycle. This study focused on the subcritical CO2 power cycle with low heat source temperature close to the critical points.

2 Methods

The water and some organic materials are working fluids used in the power plants using steam and organic Rankine cycles, respectively. In contrast, in the CO2 power cycle, carbon dioxide is used as a working fluid. The subcritical CO2 power cycle operates under temperature and pressure not exceeding the critical point for CO2. Therefore, the low heat source temperature, like ambient temperature (Tamb) or geothermal, is sufficient for the subcritical CO2 power cycle. The ideal cycle was used in this study that was an isobaric process at the heat exchangers and isentropic steps at the expansion and compression.

2.1 Components and processes of CO2 power cycle

The subcritical CO2 cycle is very similar to the simple steam Rankine cycle; therefore it is called the CO2 Rankine cycle. Evaporator, turbine, condenser, and pump are demand to configuration the CO2 power cycle, as it is shown in Fig. 1a. The T–s diagram shows the processes of the CO2 power cycle in Fig. 1b. The CO2 compressed from point 1 to point 2 by a pump in an isentropic process. A slight increase in temperature occurs, together with the rise of the pressure. Then, CO2 enters the heat exchanger (sometimes called evaporator) at point 2; here, in the initial part, the temperature will increase. Then, after reaching the boiling temperature for the given pressure, evaporation happens, even the total mass of the fluid reaches a vapor state (point 3). At point 3, the fluid is in high pressure and high temperature saturated vapor state. Here, the working fluid enters the turbine (or expander) with high pressure and temperature to produce the mechanical work by expansion between points 3 and 4. During this process (taken as ideal adiabatic, i.e., isentropic one) pressure and temperature of the working fluid is decreased. In stage 4, the fluid is in a low-enthalpy, low-temperature, low-pressure saturated vapor state. Between points 4 and 1, part of the heat is removed from the system isobarically in a second heat exchanger, called condenser, causing complete condensation from saturated steam to saturated liquid state in order to start a new cycle.

Fig. 1.
Fig. 1.

Schematic diagram for CO2 power cycle, a) Main components of cycle, b) Ts diagram for CO2 power cycle, and c) Ts diagram for TFC cycle

Citation: Pollack Periodica 16, 2; 10.1556/606.2021.00310

There is a slightly different cycle, called TFC; it is characterized by its simplicity. Heat addition terminates, when the saturated liquid state reached (point 3, Fig. 2). In this way, the fluid volume between points 2 and 3 is not significantly different, and therefore the heat exchanged for TFC can be simpler, than for ORC. In general, the process 1–2–3–4 is similar to the CO2 cycle used in the study. The two differences are the lack of evaporation in the ‘evaporator’ and the “wet expansion” (i.e., expansion starting from the liquid state) in the expander [23]. The difference can be seen in detail in Fig. 1b and c.

Fig. 2.
Fig. 2.

T–s diagram for carbon dioxide as wet working fluid with characteristic points (see in text)

Citation: Pollack Periodica 16, 2; 10.1556/606.2021.00310

2.2 Characteristics of CO2 as working fluid

Due to the thermal stability, the CO2 is one of the natural working fluids considered suitable for the refrigeration and power cycles, also; it has significant advantages compared to other working fluids that lack whole or part of these characteristics, which include physical, chemical, environmental, and economic features. There are some crucial factors that must be considered when selecting a working fluid, or it characterizes the working fluids, for chemically factors like non-flammability, low toxicity; physical factors, like high critical pressure (Pcr), low critical temperature, low boiling point (Bp), and good heat transfer; the environmental ones, like low Global Warming Potential (GWP), low Ozone-Depleting Potential (ODP), and environmentally friendliness (very safe to use); and finally, low cost as the economic factor. Carbon dioxide is one of the liquids whose properties satisfy the above-mentioned characteristics; additionally, it is a “natural” fluid. The CO2 properties and American Society of Heating, Refrigerating and Air-Conditioning Engineers (ASHRAE) for CO2, are shown in Table 1. The CO2 can be used in all types of CO2 cycles like subcritical, transcritical, supercritical; also, it can be used as pure fluid as well as in mixtures with other working fluids especially with hydrocarbons, especially because of its ability to reduce the flammability of hydrocarbons while preserving all the other desired thermodynamic properties [24]. According to the traditional working fluid classification, CO2 is a wet working fluid; in the novel classification, it belongs to the so-called ACZ class [25].

Table 1.

Properties of carbon dioxide (Source: on the basis of [26])

TypeCategoryASHRAE NO.FormulaASHRAE level for safetyODPGWPCritical temperature [K]Critical Pressure [MPa]Boiling Point [K]
WetACZ744CO2A101304.12827.3773194.75

2.3 Thermodynamic analysis

A calculation was performed to find the efficiency of the CO2 power cycle with the temperature variation of the heat source by using the MATLAB software and data from the NIST webbook [27]. On the Ts diagram of CO2, A and Z marks the minimum temperature for liquid and vapor phases, and C is the critical point, as shown in Fig. 2. The blue solid line represents the saturated liquid state, and the red dashed line represents the saturated vapor state. General equations were used. The efficiency was calculated by
η=(h3h4)(h2h1)(h3h2) ,
where the (h3h4), (h2h1), and (h3h2), are the difference in enthalpy at the turbine, pump, and evaporator respectively, and the enthalpy and other data values as entropy, pressure, and dryness fraction found by NIST webbook.

In this study, some assumptions and parameters applied like the steady-state of the operation cycle, and ignored the pressure losses during the flow. At the saturated vapor phase, the working fluid entering the turbine, and the inlet pressure changes based on the increasing the heat source temperature, while in the TFC cycle, the working fluids enter the turbine at the saturated liquid. Depended on the condenser temperature and dryness fraction, the working fluid enters the condenser then leaves it at the saturated liquid. Ten thousand steps, using different temperature pairs were used in this study (with 1,000 readings for each curve, 500 for the CO2 cycle, and 500 for the TFC cycle). The condenser line, green solid line is at low temperature, for example, 217 K, and all the horizontal lines above the condenser line, green solid line are the evaporator lines (increase with increasing heat source temperature starting from orange dot line to dark blue long dash line), up to the critical point. First, the 1–2–3–4 cycle efficiency was calculated then evaporation temperature was shifted to a higher value, while the condensation temperature was kept and the efficiency of the new cycle (1–2a–3a–3b) was determined. The process continued upwards, to the vicinity of the critical temperature (see cycle 1–2d–3d–4d) as it is shown in Fig. 3. In the next step, a new (increased) condenser temperature was taken (238 K instead of 217 K green solid line shifts upward). For this new value, efficiencies related to changing evaporation temperature (orange dot, gold dash, purple long dash dot, blue long dash dot dot, and dark blue long dash lines), were also calculated. In the following steps, the process was repeated with new condenser temperatures. Finally, obtained the whole set of efficiency values for various condensation and evaporation temperature pairs from 217 K to the critical temperature was obtained. The same processes were applied for the TFC cycle, except that the entering parameter in the turbine would be saturated liquid, as it is shown in Fig. 4. All boundary conditions are shown in Table 2.

Fig. 3.
Fig. 3.

The procedure of the calculation on a T–s diagram for CO2 power cycle

Citation: Pollack Periodica 16, 2; 10.1556/606.2021.00310

Fig. 4.
Fig. 4.

The procedure of the calculation on a T–s diagram for TFC power cycle

Citation: Pollack Periodica 16, 2; 10.1556/606.2021.00310

Table 2.

The boundary conditions

RunHeat source temperature [K]Heat sink temperature [K]RunHeat source temperature [K]Heat sink temperature[K]
1217 to Tcr2176278 to Tcr278
2238 to Tcr2387284 to Tcr284
3252 to Tcr2528290 to Tcr290
4266 to Tcr2669296 to Tcr296
5272 to Tcr27210302 to Tcr302

3 Results and discussion

Carbon dioxide can be applied as working fluid in refrigeration cycles as well as in and power cycles. This study focused on the sub-critical CO2 power cycle by utilizing low-temperature heat sources for example ambient temperature or geothermal one with several of condenser temperatures. In general, the efficiency increases with increasing the heat source temperature. It has been found in sub-critical CO2 cycles (and can be generalized to all cycles having similar, ACZ-type T–s diagram) that while this increase is usually true, choosing the maximal cycle temperature close to the critical point, inverse dependency can be seen a narrow, but definitely non-zero temperature range. In contrary, for TFC, the maximum efficiency was at the critical temperature as it is shown in Fig. 5; the upper curves represent the efficiency values of CO2 power cycles and the lower curves for TFC cycles. For example, the first curve for CO2 and TFC, the range of evaporator temperature (heat source temperature) between 217 and 304.1282 K, calculated point by point with an increment equal to (total temperature range)/500, with fixed condenser temperature 217 K. For the next curve, the condenser temperature was increased, and the calculation was repeated.

Fig. 5.
Fig. 5.

Efficiency variation with heat source temperature for CO2 and TFC cycle

Citation: Pollack Periodica 16, 2; 10.1556/606.2021.00310

Separate lines in Fig. 6 represent the variation of efficiency for CO2 power cycle with increasing heat source temperature and with fixed condenser temperature (represented by the lowest temperature value on each curve), the efficiency increases for a while with increasing heat source temperature, but then a maximum appears, close, but definitely below the critical temperature, followed by an efficiency decrease, supported by results obtained with other working fluids for ORC and TFC [28]. The red dots (diamond marker) show the positions of the efficiency maximum, which is always below the critical temperature (located at the end of the curves), also the dots showing that maximum efficiency shifted close to the critical temperature with a minimum value of efficiency in the and the maximum condenser temperature 302 K. However, even this new maximum is definitely below (although closer, than the previous maximum) the critical point. It is approaching the critical temperature and its value is very small; 0.066, located at 304.10 K (while the critical temperature is 304.1282 K). Figure 7 shows efficiency for the CO2 cycle and TFC cycle in the range where this maximum appears, up to the critical temperature. The curve represents the case with condenser temperature located at 217 K, and showing that the two efficiency curves meet at one point at the critical temperature. Figure 8 shows the efficiency values at the maximum, together with the temperature of this maximum and how it decreases with increasing condenser temperature.

Fig. 6.
Fig. 6.

Maximum efficiency (red spots) for CO2 power cycle

Citation: Pollack Periodica 16, 2; 10.1556/606.2021.00310

Fig. 7.
Fig. 7.

Efficiency for CO2 and TFC at Condenser temperature 217 K

Citation: Pollack Periodica 16, 2; 10.1556/606.2021.00310

Fig. 8.
Fig. 8.

The shift of the efficiency-maxima

Citation: Pollack Periodica 16, 2; 10.1556/606.2021.00310

4 Conclusion

The CO2 power cycle has reasonably good efficiency operating with low-temperature heat sources. It uses CO2 as a working fluid, with suitable physical, chemical, environmental, and economic characteristics, compared to other working fluids. It has been known to researchers and engineers that for thermodynamic cycles the efficiency increases with the increase in the source temperature. This study showed that the efficiency of the cycle does not always increase with increasing maximal cycle temperature (and heat source temperature); close to the critical temperature; on the contrary, it reduces efficiency. It means that an efficiency maximum can exist, and the existence of this maximum should be considered upon designing subcritical CO2 Rankine cycles with maximal cycle temperature close to the critical one. In some cases (like with solar heat), the increase of the heat source temperature goes together with the increase in installation costs. This cost increase is justified only when it is associated with proper efficiency increase; in the vicinity of the critical temperature, it is not justified. The maximum of the efficiency goes closer to the critical temperature as the condenser temperature increased, while its absolute value decreases. Also, it has been shown that the efficiency of the subcritical CO2 power cycle higher than the efficiency of TFC and their efficiency equal at the critical point.

Acknowledgements

This work was performed in the frame of the FIEK_16-1-2016-0007 project, implemented with the support provided from the National Research, Development and Innovation Fund of Hungary, financed under the FIEK_16 funding scheme. Part of the research reported in this paper and carried out at BME has been supported by the NRDI Fund (TKP2020 NC, Grant No. BME-NC) based on the charter of bolster issued by the NRDI Office under the auspices of the Ministry for Innovation and Technology.

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  • [1]

    E. Cayer, N. Galanis, M. Desilets, H. Nesreddine, and P. Roy, “Analysis of a carbon dioxide transcritical power cycle using a low temperature source,” Appl. Energ., vol. 86, no. 7–8, pp. 10551063, 2009.

    • Crossref
    • Search Google Scholar
    • Export Citation
  • [2]

    Y. T. Ge, L. Li, X. Luo, and S. A. Tassou, “Performance evaluation of a low-grade power generation system with CO2 transcritical power cycles,” Appl. Energ., vol. 227, pp. 220230, 2018.

    • Crossref
    • Search Google Scholar
    • Export Citation
  • [3]

    L. Li, T. Ge, X. Luo, and S. A. Tassou, “Experimental investigations into power generation with low grade waste heat and R245fa Organic Rankine Cycles (ORCs),” Appl. Therm. Eng., vol. 115, pp. 815824, 2017.

    • Crossref
    • Search Google Scholar
    • Export Citation
  • [4]

    C. Zamfirescu and I. Dincer, “Thermodynamic analysis of a novel ammonia-water trilateral Rankine cycle,” Thermochim. Acta, vol. 477, no. 1–2, pp. 715, 2008.

    • Crossref
    • Search Google Scholar
    • Export Citation
  • [5]

    O. Al-Oran and F. Lezsovits, “Enhance thermal efficiency of parabolic trough collector using Tungsten oxide/Syltherm 800 nanofluid,” Pollack Period., vol. 15, no. 2, pp. 187198, 2020.

    • Crossref
    • Search Google Scholar
    • Export Citation
  • [6]

    S. Mondal and S. De, “Transcritical CO2 power cycle - Effects of regenerative heating using turbine bleed gas at intermediate pressure,” Energy, vol. 87, pp. 95103, 2015.

    • Crossref
    • Search Google Scholar
    • Export Citation
  • [7]

    C. Tonkó, “Criterion for the discharge of geothermal waste water into surface water sources in Hungary,” Pollack Period., vol. 7, no. 2, pp. 129138, 2012.

    • Crossref
    • Search Google Scholar
    • Export Citation
  • [8]

    X. Lei, R. Peng, Z. Guo, H. Li, K. Ali, and X. Zhou, “Experimental comparison of the heat transfer of carbon dioxide under subcritical and supercritical pressures,” Int. J. Heat Mass Transf., vol. 152, Paper no. 119562, 2020.

    • Crossref
    • Search Google Scholar
    • Export Citation
  • [9]

    H. Tuo, “Analysis of a reheat carbon dioxide transcritical power cycle using a low temperature heat source,” in ASME 2011 International Conference on Environmental Engineering and Applications, Parts A and B, Denver, Colorado, USA, November 11–17, 2011, 2011, pp. 219225.

    • Search Google Scholar
    • Export Citation
  • [10]

    H. Chen, D. Y. Goswami, and E. K. Stefanakos, “A review of thermodynamic cycles and working fluids for the conversion of low-grade heat,” Renew. Sustain. Energ. Rev., vol. 14, no. 9, pp. 30593067, 2010.

    • Crossref
    • Search Google Scholar
    • Export Citation
  • [11]

    P. Dutta and P. Kumar, “Supercritical carbon dioxide-based power cycles,” in Encyclopedia of Sustainable Technologies, M. A. Abraham (ed.) Elsevier, pp. 419428, 2017.

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Editor(s)-in-Chief: Iványi, Amália

Editor(s)-in-Chief: Iványi, Péter

 

Scientific Secretary

Miklós M. Iványi

Editorial Board

  • Bálint Bachmann (Institute of Architecture, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Jeno Balogh (Department of Civil Engineering Technology, Metropolitan State University of Denver, Denver, Colorado, USA)
  • Radu Bancila (Department of Geotechnical Engineering and Terrestrial Communications Ways, Faculty of Civil Engineering and Architecture, “Politehnica” University Timisoara, Romania)
  • Charalambos C. Baniotopolous (Department of Civil Engineering, Chair of Sustainable Energy Systems, Director of Resilience Centre, School of Engineering, University of Birmingham, U.K.)
  • Oszkar Biro (Graz University of Technology, Institute of Fundamentals and Theory in Electrical Engineering, Austria)
  • Ágnes Borsos (Institute of Architecture, Department of Interior, Applied and Creative Design, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Matteo Bruggi (Dipartimento di Ingegneria Civile e Ambientale, Politecnico di Milano, Italy)
  • Ján Bujňák (Department of Structures and Bridges, Faculty of Civil Engineering, University of Žilina, Slovakia)
  • Anikó Borbála Csébfalvi (Department of Civil Engineering, Institute of Smart Technology and Engineering, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Mirjana S. Devetaković (Faculty of Architecture, University of Belgrade, Serbia)
  • Szabolcs Fischer (Department of Transport Infrastructure and Water Resources Engineering, Faculty of Architerture, Civil Engineering and Transport Sciences Széchenyi István University, Győr, Hungary)
  • Radomir Folic (Department of Civil Engineering, Faculty of Technical Sciences, University of Novi Sad Serbia)
  • Jana Frankovská (Department of Geotechnics, Faculty of Civil Engineering, Slovak University of Technology in Bratislava, Slovakia)
  • János Gyergyák (Department of Architecture and Urban Planning, Institute of Architecture, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Kay Hameyer (Chair in Electromagnetic Energy Conversion, Institute of Electrical Machines, Faculty of Electrical Engineering and Information Technology, RWTH Aachen University, Germany)
  • Elena Helerea (Dept. of Electrical Engineering and Applied Physics, Faculty of Electrical Engineering and Computer Science, Transilvania University of Brasov, Romania)
  • Ákos Hutter (Department of Architecture and Urban Planning, Institute of Architecture, Faculty of Engineering and Information Technolgy, University of Pécs, Hungary)
  • Károly Jármai (Institute of Energy and Chemical Machinery, Faculty of Mechanical Engineering and Informatics, University of Miskolc, Hungary)
  • Teuta Jashari-Kajtazi (Department of Architecture, Faculty of Civil Engineering and Architecture, University of Prishtina, Kosovo)
  • Róbert Kersner (Department of Technical Informatics, Institute of Information and Electrical Technology, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Rita Kiss  (Biomechanical Cooperation Center, Faculty of Mechanical Engineering, Budapest University of Technology and Economics, Budapest, Hungary)
  • István Kistelegdi  (Department of Building Structures and Energy Design, Institute of Architecture, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Stanislav Kmeť (President of University Science Park TECHNICOM, Technical University of Kosice, Slovakia)
  • Imre Kocsis  (Department of Basic Engineering Research, Faculty of Engineering, University of Debrecen, Hungary)
  • László T. Kóczy (Department of Information Sciences, Faculty of Mechanical Engineering, Informatics and Electrical Engineering, University of Győr, Hungary)
  • Dražan Kozak (Faculty of Mechanical Engineering, Josip Juraj Strossmayer University of Osijek, Croatia)
  • György L. Kovács (Department of Technical Informatics, Institute of Information and Electrical Technology, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Balázs Géza Kövesdi (Department of Structural Engineering, Faculty of Civil Engineering, Budapest University of Engineering and Economics, Budapest, Hungary)
  • Tomáš Krejčí (Department of Mechanics, Faculty of Civil Engineering, Czech Technical University in Prague, Czech Republic)
  • Jaroslav Kruis (Department of Mechanics, Faculty of Civil Engineering, Czech Technical University in Prague, Czech Republic)
  • Miklós Kuczmann (Department of Automations, Faculty of Mechanical Engineering, Informatics and Electrical Engineering, Széchenyi István University, Győr, Hungary)
  • Tibor Kukai (Department of Engineering Studies, Institute of Smart Technology and Engineering, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Maria Jesus Lamela-Rey (Departamento de Construcción e Ingeniería de Fabricación, University of Oviedo, Spain)
  • János Lógó  (Department of Structural Mechanics, Faculty of Civil Engineering, Budapest University of Technology and Economics, Hungary)
  • Carmen Mihaela Lungoci (Faculty of Electrical Engineering and Computer Science, Universitatea Transilvania Brasov, Romania)
  • Frédéric Magoulés (Department of Mathematics and Informatics for Complex Systems, Centrale Supélec, Université Paris Saclay, France)
  • Gabriella Medvegy (Department of Interior, Applied and Creative Design, Institute of Architecture, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Tamás Molnár (Department of Visual Studies, Institute of Architecture, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Ferenc Orbán (Department of Mechanical Engineering, Institute of Smart Technology and Engineering, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Zoltán Orbán (Department of Civil Engineering, Institute of Smart Technology and Engineering, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Dmitrii Rachinskii (Department of Mathematical Sciences, The University of Texas at Dallas, Texas, USA)
  • Chro Radha (Chro Ali Hamaradha) (Sulaimani Polytechnic University, Technical College of Engineering, Department of City Planning, Kurdistan Region, Iraq)
  • Maurizio Repetto (Department of Energy “Galileo Ferraris”, Politecnico di Torino, Italy)
  • Zoltán Sári (Department of Technical Informatics, Institute of Information and Electrical Technology, Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Grzegorz Sierpiński (Department of Transport Systems and Traffic Engineering, Faculty of Transport, Silesian University of Technology, Katowice, Poland)
  • Zoltán Siménfalvi (Institute of Energy and Chemical Machinery, Faculty of Mechanical Engineering and Informatics, University of Miskolc, Hungary)
  • Andrej Šoltész (Department of Hydrology, Faculty of Civil Engineering, Slovak University of Technology in Bratislava, Slovakia)
  • Zsolt Szabó (Faculty of Information Technology and Bionics, Pázmány Péter Catholic University, Hungary)
  • Mykola Sysyn (Chair of Planning and Design of Railway Infrastructure, Institute of Railway Systems and Public Transport, Technical University of Dresden, Germany)
  • András Timár (Faculty of Engineering and Information Technology, University of Pécs, Hungary)
  • Barry H. V. Topping (Heriot-Watt University, UK, Faculty of Engineering and Information Technology, University of Pécs, Hungary)

POLLACK PERIODICA
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2020  
Scimago
H-index
11
Scimago
Journal Rank
0,257
Scimago
Quartile Score
Civil and Structural Engineering Q3
Computer Science Applications Q3
Materials Science (miscellaneous) Q3
Modeling and Simulation Q3
Software Q3
Scopus
Cite Score
340/243=1,4
Scopus
Cite Score Rank
Civil and Structural Engineering 219/318 (Q3)
Computer Science Applications 487/693 (Q3)
General Materials Science 316/455 (Q3)
Modeling and Simulation 217/290 (Q4)
Software 307/389 (Q4)
Scopus
SNIP
1,09
Scopus
Cites
321
Scopus
Documents
67
Days from submission to acceptance 136
Days from acceptance to publication 239
Acceptance
Rate
48%

 

2019  
Scimago
H-index
10
Scimago
Journal Rank
0,262
Scimago
Quartile Score
Civil and Structural Engineering Q3
Computer Science Applications Q3
Materials Science (miscellaneous) Q3
Modeling and Simulation Q3
Software Q3
Scopus
Cite Score
269/220=1,2
Scopus
Cite Score Rank
Civil and Structural Engineering 206/310 (Q3)
Computer Science Applications 445/636 (Q3)
General Materials Science 295/460 (Q3)
Modeling and Simulation 212/274 (Q4)
Software 304/373 (Q4)
Scopus
SNIP
0,933
Scopus
Cites
290
Scopus
Documents
68
Acceptance
Rate
67%

 

Pollack Periodica
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Pollack Periodica
Language English
Size B5
Year of
Foundation
2006
Publication
Programme
2021 Volume 16
Volumes
per Year
1
Issues
per Year
3
Founder Akadémiai Kiadó
Founder's
Address
H-1117 Budapest, Hungary 1516 Budapest, PO Box 245.
Publisher Akadémiai Kiadó
Publisher's
Address
H-1117 Budapest, Hungary 1516 Budapest, PO Box 245.
Responsible
Publisher
Chief Executive Officer, Akadémiai Kiadó
ISSN 1788-1994 (Print)
ISSN 1788-3911 (Online)

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