## 2 Experimental Setup and Configuration
### Data Reduction  
Table 4 shows the uncertainties of the experimental parameters and measured data. Thermocouples with an accuracy of \(\pm\)0.1 \({}^{\circ}\)C were used to measure the refrigerant-side and air-side temperatures, and digital pressure gauges with an accuracy of \(\pm\)0.1% were installed to measure the refrigerant-side pressure. The mass flow rate of the Coriolis type flow meter was measured with an accuracy of \(\pm\)0.2% of the reading. A data logger of a digital power meter type with an accuracy of \(\pm\)0.5% was used to measure the electric power consumed by an electric motor used for driving two compressors. The precision limits and bias limits of all the parameters associated with cooling capacity and COP were estimated in uncertainty analysis. The average uncertainties of the experimental data on cooling capacity and the COP were determined to be 3.5% and 4.7%, respectively, using the method suggested by [PERSON] [15].  
Based on the measured data, the heat transfer rate of the refrigerant-side was calculated by Equation (1) using the refrigerant enthalpy method (ASHRAE Standard 116, 1983) [16]. The cooling capacity of the evaporator part was calculated by Equation (2) using both the air flow rate and temperature difference suggested by [PERSON] et al. (2009) [17]. The air-side cooling capacity of the evaporator part was validated against the refrigerant-side heat transfer rate of the evaporator part. The cooling capacity of the air-side was consistent with the heat transfer rate of the refrigerant-side within \(\pm\)5.0%, so the present experimental setup was found to be appropriate.  
\[\dot{Q}_{ref}=\dot{m}_{ref}\Delta h_{ref} \tag{1}\]  
\[\dot{Q}_{evap}=C_{air}\dot{V}_{air}\rho_{air}(\mathrm{T}_{\mathrm{air,out}}- \mathrm{T}_{\mathrm{air,in}}) \tag{2}\]  
The compressor work was calculated by Equation (3) using the refrigerant enthalpy method. The COP (coefficient of performance) of the tested air conditioning system was calculated by Equation (4) using the air-side cooling capacity of the evaporator part and the compressor work.  
\[W_{comp}=\dot{m}_{ref}(h_{comp,out}-h_{comp,in}) \tag{3}\]  
\begin{table}
\begin{tabular}{c c} \hline \hline
**Components** & **Conditions** \\ \hline Thermocouples (T-type) & \(\pm\)0.1 \({}^{\circ}\)C \\ Pressure gage (Sensors, PI3H) & \(\pm\)0.1\%, Max 25 MPa \\ Mass flow rate (Coriolis type) & \(\pm\)0.2\%, Max 650 kg/h \\ Data logger (E. Gate IP (V3), 2.93 W @ 12.06 V) & \(\pm\)0.5\% \\ COP & 4.7\% \\ Cooling capacity & 3.5\% \\ \hline \hline \end{tabular}
\end{table}
Table 4: Uncertainties of the experimental parameters and measured data. COP: coefficient of performance.  
\begin{table}
\begin{tabular}{c c} \hline \hline
**Regions** & **Temperature (\({}^{\circ}\)C)** \\ \hline Vietnam & 34.3 \\ Thailand & 38.4 \\ Myanmar & 38.4 \\ India & 41.9 \\ Saudi Arabia & 43.2 \\ Yemen & 43.1 \\ Average temperature & 40.0 \\ \hline \hline \end{tabular}
\end{table}
Table 3: Average outdoor temperature of the Southeast Asian region.  
\[\text{COP}=\frac{Q_{evap}}{W_{comp}} \tag{4}\]  
Both the cooling capacity and the COP as the steady state performance and cooling speed of the cabin as a transient performance were evaluated, and the cooldown speed to reach a target temperature for the cabin cooling was considered. In particular, the cooldown performances of the suggested air conditioning system for a special purpose vehicle used in Southeast Asian nations are significant parameters of the heat pump because better and quicker cooldown performance in a special purpose vehicle operated in hot weather conditions will allow the passengers with equipped clothing to feel comfortable. Moreover, transient performance can be actively and quickly controlled to meet the cooling loads in the cabin because of the two parallel refrigeration cycles with two compressors used in the air conditioning system.