###### Abstract
(2004) designed axial flow compressors for a 300 MWe SCO\({}_{2}\) power cycle with Concepts NREC-AXIAL\({}^{\text{TM}}\) and predicted efficiency as high as that of large mass flow rates [5], as shown in Table 1. [PERSON] with his co-workers (2006) designed centrifugal compressors for the same power level using Balje diagram [6], as displayed in Table 2. But neither of them did a 3D calculation for their design results. There are also many other universities and research institutes that have carried out relevant research on the design strategy of SCO\({}_{2}\) compressors [7; 8; 9; 10; 11; 12] and some research institutes have carried out experiments [13; 14; 15]. Figure 1: Application fields of the supercritical carbon dioxide Brayton cycle. More recently, [PERSON] (2014) and colleagues utilized an in-house fluid dynamic solver and simulated the test compressor of Sandia National Laboratories (SNL) based on the steady state method [16]. However, their numerical simulation was not in good agreement with SNL's experimental results. They thought this difference was caused by the limitation of CFD and the uncertainty of the experimental data. In 2014, [PERSON] and his colleagues applied the [PERSON]'s one-dimensional model to design a compressor of the 10 MW central receiver type for solar plants [17]. Results of their 1-D model are in good agreement with the simulation of Fluent software. In their calculation, the realizable \(k-\varepsilon\) turbulence model was employed. [PERSON] (2014) carried out a comparison between CFD and experimental results [18]. It was found that when the inlet parameter was far away from the critical point, CFD calculation could get fairly good results. However, when the inlet condition was approaching the critical point, their deviation became more and more obvious. [PERSON] et al. (2017) carried out steady and unsteady numerical simulations of a SCO\({}_{2}\) centrifugal compressor with ANSYS-CFX [19], which showed that the difference between the two results is small. But, it should be pointed out that the equation of state used in their calculation is not accurate enough for SCO\({}_{2}\). Figures 2 and 3 are the test components of SNL and Institute of Applied Energy (IAE), respectively. These experimental results have verified the feasibility of the one-dimensional design method of the SCO\({}_{2}\) compressor and pointed out that windage loss is an important cause of low cycle efficiency. However, according to the authors' knowledge, the influences of compressor type and working parameters on design results that were obtained by conventional design methods for air have not been studied so far. The accuracy of loss models for SCO\({}_{2}\) compressors has been rarely researched, but its importance is obvious. \begin{table}
\begin{tabular}{c c c c c c} \hline \hline
**Compressor** & **Efficiency (\%)** & \
\begin{tabular}{c} **Mass Flow Rate** \\ **(kg/s)** \\ \end{tabular} & **Pressure Ratio** &
\begin{tabular}{c} **Rotating Speed** \\ **(RPM)** \\ \end{tabular} & **Stages** \\ \hline Main compressor & 92.94 & 2574 & 2.2 & 3600 & 4 \\ Re-compressor & 93 & 1175 & 2.2 & 3600 & 8 \\ \hline \hline \end{tabular}
\end{table} 
Table 1: 300 MWe axial-flow SCO\({}_{2}\) compressor ([PERSON]). \begin{table}
\begin{tabular}{c c c c c c} \hline \hline
**Compressor** & **Efficiency (\%)** & \
\begin{tabular}{c} **Mass Flow Rate** \\ **(kg/s)** \\ \end{tabular} & **Pressure Ratio** &
\begin{tabular}{c} **Rotating Speed** \\ **(RPM)** \\ \end{tabular} & **Impeller** \\ \hline Main compressor & 85.07 & 1915 & 2.6 & 3600 & 0.887 \\ Re-compressor & 72.9 & 1331 & 2.6 & 3600 & 1.546 \\ \hline \hline \end{tabular}
\end{table} 
Table 2: 300 MWe single stage centrifugal SCO\({}_{2}\) compressor (Yifang Gong). Figure 2: Sandia National Laboratories (SNL) test component (USA). At present, centrifugal compressors are studied by many more research institutes than axial flow compressors are.