## 3 Results and Discussion
Table 6 presents the results taken from [2] for the advanced exergy analysis of the R134a refrigeration system, while the ones obtained from the authors' model are listed in Table 7.  
The comparison of the results from the advanced exergy analyses of the vapor-compression refrigeration machine using R134a is shown in Table 8. It could be observed that most of the values in Table 8 for the relative difference do not exceed \(\pm\)2.0%. However, the differences for \(\dot{E}_{D,CD}^{\ UN}\ \dot{E}_{D,CD}^{\ UN,EN}\), and \(\dot{E}_{D,CD}^{\ UN,EX}\) were above 7.0%. Such a high difference could be explained in the following way. The unavoidable exergy destruction depends on the pinch point temperature (\(\Delta T_{CD}\)) within the condenser. The assumed value of this temperature difference was equal to 0.5 K. For the sake of simplicity, this value of temperature difference was defined as the difference between the temperature of R134a on the saturated liquid line for conditions of the condenser and the temperature of the heated air at the condenser outlet (see Figure 2). For such assumptions, the condenser temperature was equal to 30.5 \({}^{\circ}\)C. In [2], the condensation temperature was 30.1 \({}^{\circ}\)C. As a result, the value of unavoidable exergy destruction obtained in this work was higher than the one presented in [2]. A similar explanation could be applied to cases involving \(\dot{E}_{D,k}^{\ UN,EN}\) and \(\dot{E}_{D,k}^{\ UN,EX}\). \begin{table}
\begin{tabular}{l c c c} \hline \hline \multirow{2}{*}{**Working Fluid**} & \multicolumn{3}{c}{\(\dot{m}\) kg\(\cdot\)s\({}^{-1}\)} \\ \cline{2-4}  & **Author’s Data** & **Morosuk and Tsatsaronis [2]** & \(\delta\)**, \%** \\ \hline R134a & 0.787 & 0.787 & 0.00 \\ Air (evaporator) & 9.945 & 9.942 & 0.03 \\ Air (condenser) & 14.546 & 14.540 & 0.04 \\ \hline \hline \end{tabular}
\end{table} 
Table 3: Mass flow rates of working fluids for the real cycle with R134a. \begin{table}
\begin{tabular}{l c c c} \hline \hline \multirow{2}{*}{**Stream**} & \multicolumn{3}{c}{\(\boldsymbol{T_{\text{ }}}\,^{\circ}\)C} \\ \cline{2-4}  & **Author’s Data** & **[PERSON] [2]** & \(\delta\)**, \%** \\ \hline Evaporator & \(-15.00\) & \(-15.00\) & 0.00 \\ temperature & 30.00 & 29.60 & 1.35 \\ \hline \hline \end{tabular}
\end{table} 
Table 4: Temperatures of evaporation and condensation for the theoretical refrigeration cycle. \begin{table}
\begin{tabular}{l c c c} \hline \hline \multirow{2}{*}{**Stream**} & \multicolumn{3}{c}{\(\boldsymbol{T_{\text{ }}}\,^{\circ}\)C} \\ \cline{2-4}  & **Author’s Data** & **[PERSON] and [PERSON] [2]** & \(\delta\)**, \%** \\ \hline Evaporator & \(-15.00\) & \(-15.00\) & 0.00 \\ temperature & 30.00 & 29.60 & 1.35 \\ \hline \hline \end{tabular}
\end{table} 
Table 5: Temperatures of evaporation and condensation for the refrigeration cycle with unavoidable exergy destruction. Generally, it could be concluded that the results of the advanced exergy analysis for the vapor-compression refrigeration machine using R134a obtained by the authors and presented in [2] are quite similar. The results obtained from the existing advanced exergy analysis with the ones evaluated on the basis of the new proposed approach are compared in Table 9. It could be concluded that \(\dot{E}_{D,k}^{AV,EN}\), \(\dot{E}_{D,k}^{AV,EX}\), \(\dot{E}_{D,k}^{AV,INT}\), and \(\dot{E}_{D,k}^{AV,EXT}\) of the evaporator and throttling valve coincided. However, large differences for the compressor and condenser were observed. According to the proposed approach, the value of internally caused exergy destruction within the compressor (\(\dot{E}_{D,CM}^{AV,INT}\)) due to irreversibilities taking place in this component was equal to 6.507 kW, which is larger by 70.4% compared to the avoidable