## 2 Case Setup
\right) \tag{5}\]  
Inlet air velocity at the exhaust vent of the fan-coil unit and a pressure outlet boundary condition is set on the outlet (located at the bottom of the fan-coil unit). The radiation heat transfer was considered by applying the surface-to-surface radiation model ([PERSON], 1992) [20]. In this model, the air is considered a transparent medium, and all surfaces are assumed to be diffusive and grey. According to [18], the following values were considered for the emissivity of surfaces: 0.88, 0.89, 0.8, and 0.9 for the fan-coil box, window (glass), and walls, respectively. Boundary conditions of the third kind were applied at both the external walls and windows. The heat losses through the external walls and windows were estimated by considering thermal transmittance (U-values) equal to 1.4 and 5.7 W/(m\({}^{2}\)K), respectively, and a convection coefficient equal to 25 W/(m\({}^{2}\)K), while the remaining walls and the door were considered isothermal. Furthermore, the floor and ceiling surfaces were considered adiabatic. The summer-cooling simulations were performed for a typical day of July when the outside temperature is 30 \({}^{\circ}\)C, while the isothermal internal walls are at a static temperature of 24 \({}^{\circ}\)C. Four different simulations were performed, varying the inlet air mass flow rate and the presence of the solar load. The values of the direct solar radiation (597.6 W/m\({}^{2}\) with a sunshine factor of 1) and diffuse solar radiation (597.6 W/m\({}^{2}\) with a diffuse fraction of 0.5) for 15 July 2021 were considered according to the solar-load tool of StarCCM+. When the solar load was disregarded, the direct radiation of the sun was set to 0, while low diffuse radiation of about 10 W/m\({}^{2}\) was considered due to the presence of a highly reflective adjacent building. The considered cases in this study are summarised in Table 1.  
For the winter-heating simulations, 6 February 2021 was considered, when an external temperature of 6 \({}^{\circ}\)C was recorded. The isothermal internal walls were considered at a static temperature of 21\({}^{\circ}\)C. The two studied cases only differed for the input air velocity and are summarised in Table 2. \begin{table}
\begin{tabular}{c c c c} \hline  & **Season** & **Inlet Air Velocity** & **Solar Load** \\  & & **(m/s)** & **Solar Load** \\ \hline Case A & Cooling & 1 & NO \\ Case B & Cooling & 2 & NO \\ Case C & Cooling & 1 & YES \\ Case D & Cooling & 2 & YES \\ \hline \end{tabular}
\end{table} 
Table 1: Summary of the studied configurations in summer-cooling conditions. Mesh sensitivity analysis was performed to obtain the best scenario in terms of computational cost and accuracy of the results. Five different values of the base size of the mesh were considered. Velocity and temperature values in different positions inside the room were analysed. Figure 2 shows that the fourth mesh was the best choice in terms of trade-off between computational time and solution accuracy. Thus, the final mesh consisted of \(2\times 10^{6}\) polyhedral elements with a base size of 6 cm. A nonuniform mesh strategy was regarded with a denser mesh in the vicinity of solid surfaces and fan-coil vents, and smoother mesh density with an expansion rate for far-field regions. A mesh refinement was performed on the furniture (4 cm) and chair (2 cm) surfaces due to their small geometrical dimensions with respect to the overall domain. In addition, a boundary layer was modelled in the fluid zone with five layers with a total thickness of 5% of the base size, resulting in a maximal y+ value of 3.5. An overview of the final mesh can be seen in Figure 3.  
The local parameters obtained with the simulations were used to evaluate the levels of discomfort inside the office with a focus on the working stations to define which the unfavourable ones were.