Table 2 shows the air properties as a function of the air inlet temperature at the evaporator.

The air passage area inside the heat exchanger is represented by Equation (25).

$Ase{c}_{NHp}=AEn{t}_{HE}-\left(N_{HP}\ast \pi \ast \frac{D_{ext}^2}{4.0}-N_{TotFin}\ast H_{Fin}\ast T_{Fin}\right)$(25)

The hydraulic perimeter of the heat exchanger is represented by Equation (26).

$P_{hHp}=N_{HP}\ast P_{HP}+N_{TotFin}\ast \left(2.0\ast H_{Fin}+T_{Fin}\right)-N_{TotFin}\ast H_{Fin}$(26)

The hydraulic diameter of the heat exchanger is represented by Equation (27).

$D_{hHP}=\frac{4.0\ast Ase{c}_{NHp}}{P_{hHp}}$(27)

The heat exchange area associated with the fins is represented by Equation (28).

$A_{trFin}=N_{Fin}\ast N_{HP}\ast L_{EV}\ast \left(2.0\ast H_{Fin}+T_{Fin}\right)$ (28)

The heat exchange area associated with the heat pipe assembly is represented by Equation (29).

$A_{trEv}=L_{EV}\ast N_{HP}\ast \left(P_{HP}-N_{Fin}\ast T_{Fin}\right)$(29)

The total heat exchange area in the evaporator is represented by Equation (30).

$A_{totEv}=A_{trFin}+A_{trEv}$(30)

The Reynolds number associated with the air flow inside the evaporator is represented by Equation (31).

$R{e}_{AirEv}=\frac{4.0\ast {\stackrel{\dot{}}{m}}_{air}}{\pi \ast D_{hHP}\ast {\mu }_{air}}$(31)

The air velocity inside the evaporator is represented by Equation (32).

$V_{AirEv}=\frac{R{e}_{AirEv}\ast {\mu }_{air}}{D_{hHP}}$(32)

The Reynolds number associated with the air flow at the evaporator inlet is represented by Equation (33).

$R{e}_{AirEnt}=\frac{4.0\ast {\stackrel{\dot{}}{m}}_{air}}{\pi \ast D_{Shell}\ast {\mu }_{air}}$(33)

The air velocity at the evaporator inlet is represented by Equation (34).

$V_{AirEnt}=\frac{R{e}_{AirEnt}\ast {\mu }_{air}}{D_{Shell}}$(34)

The saturation temperature difference across the evaporator is represented by Equation (35).

$\Delta T_{Evsat}=T_{airin}-T_{sat}$(35)

The estimated heat transfer coefficient for the boiling process, $h_{boil}$, Equation (36), was represented by Gupta and Varshney correlation [16] , depend on the $Hea{t}_{Flux}$, described by Pioro correlation [17] , Equation (37).

$\begin{array}{l}{h}_{boil}=\frac{k_{Water}}{l_{\ast }}\ast 1.39\ast {\left[Hea{t}_{Flux}\ast {\rho }_{Water}\ast C{p}_{Water}\ast \frac{l_{\ast }}{{\rho }_l\ast h_{lv}\ast k_{Water}}\right]}^{0.7}\\ \ast {\left({\rho }_l/{\rho }_v\right)}^{0.21}\ast {\left({\mu }_{Water}\ast \frac{C{p}_{Water}}{k_{Water}}\right)}^{-0.21}\end{array}$ (36)

$Hea{t}_{Flux}={\mu }_{Water}\ast h_{lv}\ast l_{\ast }\ast \left[{\left(\frac{1.0}{C_{sf}}\right)}^{0.33}\ast {\left(P{r}_{Water}\right)}^{1.0/0.33}\ast {\left(C{p}_{Water}\ast \frac{\Delta T_{Evsat}}{h_{lv}}\right)}^{1.0/0.33}\right]$(37)

where,

$l_{\ast }={\left[\frac{{\sigma }_{Water}}{g\ast \left({\rho }_l-{\rho }_v\right)}\right]}^{1.0/2.0}$(38)

The Nusselt number associated with air is represented by Equation (39).

$N{u}_{airEv}=0.696\ast {\left(R{e}_{AirEv}\right)}^{0.5}\ast {\left(P{r}_{air}\right)}^{0.36}\ast {\left(\frac{P{r}_{air}}{5.0\ast P{r}_{air}}\right)}^{0.25}$(39)

The heat transfer coefficient by air convection in the evaporator is obtained by:

$h_{Ev}=N{u}_{air}\ast \frac{k_{air}}{D_{ext}}$(40)

The application of the concept of “Aleta Analogy”, conceived by Fakheri [1] leads us to define the following parameters:

$m{L}_{EvFin}=\sqrt{\frac{2h_{Ev}}{k_{Fin}{t}_{Fin}}}{P}_{hEv}$ (41)

${\eta }_{EvFin}=\frac{\text{Tanh}\left(m{L}_{EvFin}\right)}{m{L}_{EvFin}}$ (42)

The fin efficiency for the evaporator section is defined through Equation (42) by ${\eta }_{EvFin}$.

${\beta }_{Ev}=\frac{A_{trFin}}{A_{totEv}}$(43)

${{\eta }^{\prime }}_{EvFin}={\beta }_{Ev}{\eta }_{EvFin}+\left(1-{\beta }_{Ev}\right)$ (44)

The efficiency associated with the set of fins in the evaporator, weighted by the area of change of the fins ${{\eta }^{\prime }}_{EvFin}$, is represented through Equation (44).

$U{o}_{Ev}=\frac{1}{\frac{1}{h_{boil}}+\frac{D_{ext}-D_{int}}{k_W}+\frac{1}{{{\eta }^{\prime }}_{EvFin}{h}_{Evair}}}$(45)

The global heat transfer coefficient associated with air in the evaporator, $U{o}_{Ev}$, is given by Equation (45).

$C_{Air}={\stackrel{\dot{}}{m}}_{air}C{p}_{air}$ (46)

The heat capacity of the air in the evaporator, $C_{Air}$, is given by Equation (46).

$C_{Ev}=C_{air}$ (47)

$NT{U}_{Ev}=\frac{U{o}_{Ev}\ast A_{totEv}}{C_{Ev}}$(48)

The number of thermal units associated with air in the evaporator, $NT{U}_{Ev}$, is given by Equation (48).

$F{a}_{Ev}=\frac{NT{U}_{Ev}}{2.0}\text{for}\text{parallel flow}$ (49)

The dimensionless number, called “fin analogy,” $F{a}_{Ev}$ is represented by Equation (49) as defined by Ahamad Fakheri [12] and reported by Nogueira, É. [9] - [11] .

${\eta }_{TEv}=\frac{\tanh \left(F{a}_{eV}\right)}{F{a}_{Ev}}$ (50)

The thermal efficiency associated with the evaporator is ${\eta }_{TEv}$ [9] - [11] .

${\epsilon }_{TEv}=\frac{1}{\frac{1}{{\eta }_{tEv}NT{U}_{Ev}}+\frac{1}{2}}$(51)

The thermal effectiveness associated with the heat evaporator is ${\epsilon }_{TEv}$.

${\stackrel{\dot{}}{Q}}_{Ev}=\frac{C_{Ev}\Delta T_{Evsat}}{\frac{1}{{\eta }_{TEv}NT{U}_{Ev}}+\frac{1}{2}}$ (52)

The heat transfer rate between the air and the heat pipe in the evaporating region ${\stackrel{\dot{}}{Q}}_{Ev}$ is given by Equation (52).

$T_{airout}=T_{airin}-\frac{{\stackrel{\dot{}}{Q}}_{Ev}}{C_{Ev}}$ (53)

After passing through the evaporator (precooling), the outlet air temperature is represented through Equation (53).

Thermal irreversibility in the evaporator is obtained using Equation (54), below.

$Irre{v}_{TEv}=\log \left(\frac{T_{airin}}{T_{airout}}\right)$(54)

The rate of thermal entropy generation is represented by Equation (55).

$\stackrel{\dot{}}{S}ge{n}_{TEv}=Irre{v}_{TEv}\ast C_{Ev}$ (55)

$Fri{c}_{Ev}=\frac{0.31}{{\left(R{e}_{AirHP}\right)}^{0.25}}$(56)

The friction factor in the evaporator is represented by Equation (56).

$\Delta p_{Ev}=Fri{c}_{Ev}\ast \frac{L_{Ev}}{D_{hEv}}\ast \frac{{\rho }_{Air}\ast {\left(V_{AirEv}\right)}^{2.0}}{2}$ (57)

The pressure drops across the evaporator is represented by Equation (57).

The pressure at the evaporator inlet is represented by Equation (58).

$P_{1Ev}=P_{2Ev}+\Delta p_{Ev}$(58)

where,

$P_{2Ev}=P_{atm}$(59)

Viscous irreversibility in the evaporator is obtained using Equation (60) below.

$Irre{v}_{VEv}=\log \left(\frac{P_{2Ev}}{P_{1Ev}}\right)$(60)

The rate of viscous entropy generation is represented by Equation (61).

$\stackrel{\dot{}}{S}ge{n}_{VEv}=Irre{v}_{VEv}\ast C_{Ev}$(61)

The Bejan number in the evaporator is represented by Equation (62).

$B{e}_{Ev}=\frac{\stackrel{\dot{}}{S}ge{n}_{TEv}}{\stackrel{\dot{}}{S}ge{n}_{TEv}+\stackrel{\dot{}}{S}ge{n}_{VEv}}$(62)

Table 3 shows the air properties as a function of the air inlet temperature at the condenser.

The heat exchange area of the heat pipes in the condenser is represented by Equation (63).

$A_{trCd}=L_{Cond}\ast N_{HP}\ast \left(P_{HP}-N_{Fin}\ast T_{Fin}\right)$(63)

The heat exchange area in the condenser is represented by Equation (64).

$A_{totCd}=A_{trFin}+A_{trCd}$(64)

The Reynolds number associated with the air flow inside the condenser is represented by Equation (65).

$R{e}_{AirCd}=\frac{4.0\ast {\stackrel{\dot{}}{m}}_{air}}{\pi \ast D_{hHP}\ast {\mu }_{air}}$(65)

The air velocity inside the condenser is represented by Equation (66).

$V_{AirCd}=\frac{R{e}_{AirCd}\ast {\mu }_{air}}{D_{hHP}}$ (66)

The saturation temperature difference across the condenser is represented by Equation (67).

$\Delta T_{Cdsat}=T_{sat}-T_{airin}$(67)

The condensation coefficient in the condenser is represented by Equation (68).

$h_{Cond}=0.943\ast {\left(\frac{{\rho }_l\ast \left({\rho }_l-{\rho }_v\right)\ast h_{lv}\ast g\ast {\left(k_{water}\right)}^3}{{\mu }_{awter}\ast L_{Cd}\ast \Delta T_{Cdsat}}\right)}^{\text{1}/\text{4}}$(68)

The Nusselt number associated with the air in the condenser is represented by Equation (69).

$N{u}_{airCd}=0.696\ast {\left(R{e}_{AirCd}\right)}^{0.5}\ast {\left(P{r}_{air}\right)}^{0.36}\ast {\left(\frac{P{r}_{air}}{5.0\ast P{r}_{air}}\right)}^{0.25}$ (69)

The heat convection coefficient associated with the air in the condenser is represented by Equation (70).

$h_{Cd}=N{u}_{airCd}\ast \frac{k_{air}}{D_{ext}}$(70)

The application of the concept of “Aleta Analogy”, conceived by Fakheri [1] leads us to define the following parameters:

$m{L}_{CdFin}=\sqrt{\frac{2h_{Ev}}{k_{Fin}{t}_{Fin}}}{P}_{hCd}$ (71)

${\eta }_{CdFin}=\frac{\text{Tanh}\left(m{L}_{CdFin}\right)}{m{L}_{CdFin}}$ (72)

The fin efficiency for the condenser section is defined through Equation (72) by ${\eta }_{CdFin}$.

${\beta }_{Cd}=\frac{A_{trFin}}{A_{totCd}}$(73)

${{\eta }^{\prime }}_{CdFin}={\beta }_{Cd}{\eta }_{CdFin}+\left(1-{\beta }_{Cd}\right)$(74)

The efficiency associated with the set of fins in the condenser, weighted by the area of change of the fins ${{\eta }^{\prime }}_{EvFin}$, is represented through Equation (74).

$U{o}_{Cd}=\frac{1}{\frac{1}{h_{cond}}+\frac{D_{ext}-D_{int}}{k_W}+\frac{1}{{{\eta }^{\prime }}_{CdFin}{h}_{Cd}}}$(75)

The global heat transfer coefficient associated with air in the condenser, $U{o}_{Cd}$, is given by Equation (75).

$C_{Air}={\stackrel{\dot{}}{m}}_{air}C{p}_{air}$ (76)

The heat capacity of the air in the condenser, $C_{Air}$, is given by Equation (76).

$C_{Cd}=C_{air}$ (77)

$NT{U}_{Cd}=\frac{U{o}_{Cd}\ast A_{totCd}}{C_{Cd}}$(78)

The number of thermal units associated with air in the condenser, $NT{U}_{Cd}$, is given by Equation (78).

$F{a}_{Cd}=\frac{NT{U}_{Cd}}{2.0}\text{for counter flow}$(79)

The dimensionless number, called “fin analogy,” $F{a}_{Cd}$ is represented by Equation (79) as defined by Ahamad Fakheri [1] and reported by Nogueira, É. [11] - [13] .

${\eta }_{TCd}=\frac{\tanh \left(F{a}_{Cd}\right)}{F{a}_{Cd}}$(80)

The thermal efficiency associated with the condenser is ${\eta }_{TCd}$.

${\epsilon }_{TCd}=\frac{1}{\frac{1}{{\eta }_{tCd}NT{U}_{Cd}}+\frac{1}{2}}$(81)

The thermal effectiveness associated with the condenser is ${\epsilon }_{TCd}$.

${\stackrel{\dot{}}{Q}}_{Cd}=\frac{C_{Cd}\Delta T_{Evsat}}{\frac{1}{{\eta }_{TCd}NT{U}_{Cd}}+\frac{1}{2}}$(82)

The heat transfer rate between the air and the heat pipe in the condenser region ${\stackrel{\dot{}}{Q}}_{Cd}$ is given by Equation (82).

$T_{airout}=\frac{{\stackrel{\dot{}}{Q}}_{Cd}}{C_{Cd}}+T_{airin}$(83)

After passing through the condenser (heat recover), the outlet air temperature is represented through Equation (83).

Thermal irreversibility in the condenser is obtained using Equation (84) below.

$Irre{v}_{TCd}=\log \left(\frac{T_{airin}}{T_{airout}}\right)$ (84)

The rate of thermal entropy generation in the condenser is represented by Equation (85).

$\stackrel{\dot{}}{S}ge{n}_{TCd}=Irre{v}_{TCd}\ast C_{Cd}$(85)

$Fri{c}_{Cd}=\frac{0.31}{{\left(R{e}_{AirCd}\right)}^{0.25}}$ (86)

The friction factor in the condenser is represented by Equation (86).

$\Delta p_{Cd}=Fri{c}_{Cd}\ast \frac{L_{Cd}}{D_{hCd}}\ast \frac{{\rho }_{Air}\ast {\left(V_{AirCd}\right)}^{2.0}}{2}$(87)

The pressure drops across the condenser is represented by Equation (87).

The pressure at the evaporator inlet is represented by Equation (88).

$P_{1Cd}=P_{2Cd}+\Delta p_{Cd}$(88)

By definition:

$P_{2Cd}=P_{atm}$(89)

Viscous irreversibility in the condenser is obtained using Equation (90) below.

$Irre{v}_{VCd}=\log \left(\frac{P_{2Cd}}{P_{1Cd}}\right)$(90)

The rate of viscous entropy generation in the condenser is represented by Equation (91).

$\stackrel{\dot{}}{S}ge{n}_{VCd}=Irre{v}_{VCd}\ast C_{Cd}$(91)

The Bejan number in the condenser is represented by Equation (92).

$B{e}_{Cd}=\frac{\stackrel{\dot{}}{S}ge{n}_{TCd}}{\stackrel{\dot{}}{S}ge{n}_{TCd}+\stackrel{\dot{}}{S}ge{n}_{Cd}}$(92)

This section presents results for thermal and viscous performances related to the evaporator of the axially finned heat pipe heat exchanger, for configurations with 29 and 47 heat pipes.

Figure 3 presents the average air velocities for the two configurations considered in this work, with shell diameters equal to 614 mm (29 heat pipes) and 789 mm (47 heat pipes). As expected, the inlet velocity of the heat exchanger with the larger diameter has a lower velocity.

Figure 4 shows the average velocity inside the heat exchangers considered in the evaporator analysis. As expected, due to its larger equivalent diameter, the heat exchanger with 47 heat pipes has a lower internal velocity, despite the larger area occupied by the heat pipes and fins. Figure 2 shows a cross-sectional representation of the finned heat pipes arranged symmetrically inside the shell for the 29-heat pipe configuration. The radial configuration is designed to achieve homogeneous air distribution inside the heat exchanger.

The maximum permissible velocity inside the heat exchanger is 6 m/s. This value limits the flow rate that can be used. For a 49-tube heat exchanger, this value is achieved with a mass flow rate of 0.2 kg/s. A standard 60 m³ operating room with 20 air changes per hour, as defined by country-specific standards, requires a flow rate of 0.4 kg/s. In this case, two 49-tube heat exchangers are needed to meet the pre-established requirements.

As expected, the Reynolds number in the evaporator, represented in Figure 5 , demonstrates a similar trend to the internal velocity in the evaporator, with the heat exchanger with 47 heat pipes presenting slightly lower values for higher air flows.

The Nusselt number, Figure 6 , presents results like those obtained and represented in Figure 3 , as it is strongly dependent on the Reynolds number.

Figure 7 presents values for the boiling heat transfer coefficient. The boiling coefficient presents significantly higher values with increasing air temperature, since it depends on the temperature difference between the air inlet temperature and the saturation temperature of the working fluid.

The number of thermal units in the evaporator is represented in Figure 8 , as a function of the air inlet flow rate and for the maximum temperature, equal to 45˚C. The working fluid filling ratio in the heat pipes, in purely theoretical terms, varies from 50% to 100%. The number of thermal units increases with the filling rate, and the number of heat pipes and decreases with an increasing flow rate. The increase in the number of heat pipes can be explained by the increase in the heat exchange area, and the increase in the filling rate is explained by the higher value of the overall heat exchange coefficient, which is strongly dependent on the boiling coefficient.

Figure 9 presents result for evaporator thermal effectiveness, for an air inlet temperature of 45˚C and a working fluid fill ratio ranging from 50% to 100%. As expected, thermal effectiveness presents results like those of the number of thermal units, since it is strongly dependent on this parameter. What stands out in the results obtained are the high thermal effectiveness values across the entire flow rate range under analysis. The extremely high effectiveness values, even for a working fluid fill ratio of 50%, demonstrate that the 47-finned heat pipe configuration has great potential for use in air conditioning systems for thermal comfort. Even for the 29-tube configuration, the results presented, for a fill ratio of 60%, are extremely promising.

The heat transfer rate in the evaporator, for an air inlet temperature of 45˚C, is represented in Figure 10 . The results obtained reflect what was observed for thermal effectiveness, with higher values for a greater number of heat pipes and higher working fluid filling rates. It should be noted that the analysis is restricted to the final position of the evaporator contained by the working fluid, since this is the most relevant region in terms of heat exchange.

The air outlet temperature is represented in Figure 11 . The results obtained for the 47-finned heat pipe configuration, as expected, are extremely promising, since the temperature, for a 60% fill ratio, presents outlet temperatures between 27˚C and 28˚C, for almost the entire flow rate range under analysis. Even for the 29-tube configuration, the results obtained for a fill fraction of 80%, below 30˚C and an air inlet temperature of 45˚C, demonstrate exceptional thermal performance for the theoretical design of the heat exchanger under analysis.

The current design, with axially finned heat pipes, is compared with the design of radially finned heat pipes, whose theoretical-experimental study is described through the work presented by Ragil Sukarno et al. [3] , and a comparative theoretical study, through the Thermal Efficiency Method, carried out by Élcio Nogueira [11] .

The results of the comparison performed in this work are shown in Figure 12 , for air outlet temperature as a function of flow rate. It can be observed that the axial design with 47 finned heat pipes presents better thermal performance than the radial configuration, for a fill ratio of 100%. However, for an effective comparison, the heat exchange areas of each heat exchanger design must be analyzed: For 29 axially finned heat pipes, with 25 fins per tube, the total heat exchange area in the evaporator is 0.17 m²; for 47 axially finned heat pipes, with 25 fins per tube, the total heat exchange area in the evaporator is 0.27 m²; in the case of radially finned heat pipes, the total heat exchange area in the evaporator is 0.39 m². In terms of thermal performance in the evaporator, the axially finned tube design offers better performance, as it provides superior results for a smaller heat exchange area.

The results and analyses presented for thermal performance, in the evaporator, validate and demonstrate better thermal performance for the current configuration.

A heat exchanger analysis, which takes into account the cost-benefit in terms of thermal/viscous performance, is presented below for the evaporator.

Figure 13 presents the coefficient of friction results for the two designs analyzed in this work: axially finned heat pipes with 29 tubes and 47 tubes, both with 25 fins per tube. Since the design with the greater number of tubes and fins has a larger area, it would be expected to have a significantly higher coefficient of friction. However, it can be seen that the velocity ( Figure 4 ) is lower for the greater number of tubes, since the air passage area is larger. This characteristic balances the coefficient of friction results, making the result obtained for 47 tubes slightly higher than the result obtained for 29 heat pipes. As will be discussed later, this result has a positive impact on thermal/viscous performance in the current configuration.

The results for evaporator pressure drop are shown in Figure 14 . The pressure drop for the 47-tube configuration shows a lower pressure drop than that for the 29-tube configuration. These results reflect what was observed for internal velocities in the heat exchanger, since the larger the number of tubes, the larger the air passage area.

The results for the evaporator’s Bejan number are presented in Figure 15 for the two configurations under analysis: 29 and 47 axially finned heat pipes. Initially, the significant superiority of the 47-heat pipe design can be observed, as it offers a higher Bejan number for a given air mass flow rate. In this case, the result reflects higher thermal performance and lower pressure loss for the 47-heat pipe configuration. However, what is very favorable, in terms of cost-benefit, is that the value obtained is extremely high for both cases. These results demonstrate that the heat exchanger designs analyzed are extremely promising, as they offer high heat transfer rates and low pressure drops.

This section presents results for thermal and viscous performances related to the condenser of the axially finned heat pipe heat exchanger, for configurations with 29 and 47 heat pipes.

The results obtained for the condenser are very similar to those obtained for the evaporator. Numerically, however, they present different values, since the length of the condenser is equal to half the length of the evaporator, and the temperature differences between the air inlet temperature and the saturation temperature are smaller. The air inlet temperature to the evaporator varies between 18˚C and 26˚C.

The air velocity values inside the condenser ( Figure 16 ) are lower than those observed for the evaporator, due to the condenser’s smaller hydraulic diameter. The velocity values for the 47-heat pipe configuration, which has a larger air passage area, are reflected in lower velocity values.

Figure 17 presents result for the condensation heat transfer coefficient. The values are lower than those obtained for the boiling coefficient, due to the high heat transfer resistance presented by the water vapor fraction inside the heat pipes. This result is expected to impact the condenser’s thermal performance.

Figure 18 presents the results for the condenser heat transfer rate as a function of the air mass flow rate for the two configurations under analysis. The input parameter is an air temperature of 18˚C. The condenser heat transfer rate is higher for the 47-heat pipe design due to the larger heat exchange area.

A comparison between axially finned and radially finned heat pipe designs for evaporator outlet air temperature is shown in Figure 19 . The thermal performance of the axially finned heat pipe configurations is significantly superior for both 29-tube and 47-tube configurations. The results obtained for the condenser are qualitatively different from those obtained for the evaporator. The numerical superiority of the 47-tube heat pipe configuration is explained by the larger heat exchange area, in relation of the 29-tube heat exchanger. The superiority of the 29-tube configuration, compared to the crossflow arrangement, despite the greater number of heat pipes, can be explained by the heat exchanger configuration. In the condenser, the countercurrent flow configuration offers superior thermal performance compared to the parallel flow configuration, as in the evaporator. Another relevant aspect is that the hydraulic diameter in the condenser is smaller than that of the evaporator, and this affects the external flow in relation to the heat pipes, increasing the Reynolds number.

The results presented for the Bejan number in the condenser are represented in Figure 20 . As expected, based on what was already discussed for the evaporator, the Bejan number is higher for the 47-tube configuration because it has a larger heat exchange area and air passage area. However, what is noteworthy, once again, is the thermal performance versus viscous dissipation, which is extremely high for both configurations analyzed. In this sense, the cost-benefit associated with the heat exchanger is very favorable, due to its high thermal performance.

The air outlet configurations in the heat exchanger under analysis, represented in Figures 2(a)-(c) , are highlighted in the last one, as it more functionally encompasses the first two. Despite this, it cannot yet be considered definitive for two reasons: four air outlets, despite resulting in a completely symmetrical and functional configuration, are not a practical option when directing cooled or heated air to a given installation; and to achieve the final solution, due to the air-directing curves, it was necessary to extend the length of the adiabatic region (200 mm). This length does not correspond to the length of the adiabatic region (120 mm) considered in the simulation, since it was decided to maintain the dimensions of the heat exchanger that had already demonstrated satisfactory experimental results and was the subject of comparison in this work.

If the heat exchanger outlet area is included in the heat exchanger analysis, its viscous dissipation must be added to the viscous dissipation rate determined for the evaporator and condenser lengths, which will negatively affect the overall performance of the heat exchanger. In this case, the Bejan number should decrease, since the total irreversibility, thermal + viscous, increases. To minimize this undesirable but unavoidable effect, the air outlet should be designed to minimize viscous losses as much as possible. However, expectations in this case are positive since the heat exchange performance is exceptionally high.