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Thermal performance of desiccant-based solar air-conditioning system with silica gel coating

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Abstract

In the present paper, a solar-driven air-conditioning system comprising silica gel-coated concentric tube heat exchanger is fabricated and analyzed experimentally. The setup consists of two concentric tube heat exchangers with silica gel coating to achieve continuous dehumidification of the humid air. Parabolic trough-type solar collector is used as air heater to supply hot air continuously to regenerate the silica gel. The system performance is measured in terms of dehumidification factor and cooling capacity. The performance parameters are plotted against the different values of the atmospheric air temperature, cooling water temperature, and specific humidity. It is observed that the dehumidification factor and cooling capacity depend on both the cooling water temperature and the conditions of ambient air. The maximum value of dehumidification factor and cooling capacity achieved are 11.05 g/kg and 5.748 kJ/min, respectively, for the ambient air temperature of 37.6 °C and specific humidity of 23.99 g/kg.

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Abbreviations

SCCTHE:

Silica gel-coated concentric tube heat exchanger

ΔD :

Dehumidification factor (g/kg)

W in :

Specific humidity of air at inlet of SCCTHE (g/kg)

W out :

Specific humidity of air at outlet of SCCTHE (g/kg)

Q :

Cooling capacity (kJ/min)

:

Air mass flow rate (kg/min)

h in :

Ambient air enthalpy at inlet of experimental setup (kJ/kg)

h out :

Supply air enthalpy at outlet of experimental setup (kJ/kg)

Δz :

Absolute error

Δz/z :

Relative error

SH:

Specific humidity (g/kg)

In/out:

Inlet/outlet

References

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Correspondence to Sunil Nain.

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Appendix

Appendix

The detailed uncertainty formulation of dehumidification factor and cooling capacity is given below.

Dehumidification factor (ΔD)

$${\text{Dehumidification}}\;{\text{factor}}\; (\Delta D )= W_{\text{in}} {-}W_{\text{out}} \;({\text{g/kg)}}$$
(5)

The dehumidification factor (ΔD) depends on the specific humidity of air at the inlet and outlet of SCCTHE. The specific humidity depends on the temperature and relative humidity of air which are measured by the RTD PT 100 thermocouple (± 0.2 °C) and hygrometer RHT-200 (± 2%RH) respectively. Applying the Kline and McClintock root-mean-square method (Kline and McClintock 1953) on Eq. (5), the uncertainty in the humidity ratio/specific humidity at inlet and outlet can be found out by the given expression.

$$\pm \, W = \left[ {\left( {\frac{\partial W}{\partial t} \times \in_{t} } \right)^{2} + \left( {\frac{\partial W}{\partial \phi } \times \in_{\phi } } \right)^{2} } \right]^{1/2}$$
(6)

where ± W is the uncertainty in specific humidity, \(\in_{t}\) is the error associated with RTD PT 100 temperature sensor, \(\in_{\phi }\) is the error associated with Hygrometer RHT 300.

To solve Eq. (7), following expressions are used (Slayzak and Joseph 1998)

$$\frac{\partial W}{\partial t} = 0.622\left( {\frac{{\partial p_{\text{v}} }}{\partial t}} \right)\left[ {\frac{p}{{(p - p_{\text{v}} )^{2} }}} \right]$$
(7)
$$\frac{{\partial p_{v} }}{\partial t} = p_{\text{v}} \left( {\frac{{ - C_{8} }}{{T^{2} }} + C_{10} + 2C_{11} T + 3C_{12} T^{2} + \frac{{C_{13} }}{T}} \right)$$
(8)

And

$$\frac{\partial W}{\partial \phi } = 0.622 \times p_{\text{vs}} \left[ {\frac{p}{{(p - p_{\text{v}} )^{2} }}} \right]$$
(9)

where p, total pressure (kPa); pv, partial pressure of water vapor (kPa); pvs, saturation pressure of water vapor (kPa); ϕ, relative humidity %, decimal RH; t, dry-bulb temperature (°C), T, absolute temperature (K).

The coefficients C8–C13 are found in ASHRAE Handbook of Fundamentals (1997).

Finally, the uncertainties found in the specific humidity at inlet/outlet are used to calculate the uncertainty in the dehumidification factor by using the following expression

$$\pm\, \Delta D = \left[ {\left( {\frac{\partial \Delta D}{{\partial W_{\text{in}} }} \times \delta W_{\text{in}} } \right)^{2} + \left( {\frac{\partial \Delta D}{{\partial W_{\text{out}} }} \times \delta W_{\text{out}} } \right)^{2} } \right]^{1/2}$$
(10)

where δWin and δWout are the errors associated with specific humidity of air at inlet and outlet. The solution for Eq. (10) using average experimental data is

$$\pm \,\Delta D = 0.13947^{{}}$$
$${\text{The}}\;{\text{percentage}}\;{\text{error}}\;{\text{in}}\;{\text{dehumidification}}\;{\text{factor}}\; = \;\frac{ \pm\, \Delta D}{\Delta D} \times 100 =\,\pm\, 3.9\%$$

Cooling capacity (Q)

$${\text{Cooling}}\;{\text{capacity}}\; (Q )= \dot{m}(h_{\text{in}} - \, h_{\text{out}} ) \;({\text{kJ/min)}}$$
(11)

The cooling capacity depends on the enthalpy and mass flow rate of air. The mass flow rate of air is taken as constant throughout the experimentation. The enthalpy depends on the temperature and relative humidity which are measured by the same instruments discussed above viz. RTD PT 100 thermocouple (± 0.2 °C) and hygrometer RHT-200 (± 2%RH) respectively. The uncertainty in the enthalpy at inlet and outlet can be found out by using the Kline and McClintock root-mean-square method.

$$\pm\, h = \left[ {\left( {\frac{\partial h}{\partial t} \times \in_{t} } \right)^{2} + \left( {\frac{\partial h}{\partial \phi } \times \in_{\phi } } \right)^{2} } \right]^{1/2}$$
(12)

To find out the uncertainty in enthalpy following equation (ASHRAE Handbook of Fundamentals 1997) is used.

$$h = 1.006t + W(2501 + 1.805t)$$
(13)

where

$$\frac{\partial h}{\partial t} = 1.006 + 1.805W$$
(14)

and

$$\frac{\partial h}{\partial \phi } = \frac{\partial h}{\partial W} \times \frac{\partial W}{\partial \phi }$$
$$\frac{\partial h}{\partial W} = 2501 + 1.805t$$
(15)
$$\frac{\partial W}{\partial \phi } = 0.622 \times p_{\text{vs}} \left[ {\frac{p}{{\left( {p - p_{\text{v}} } \right)^{2} }}} \right]$$
(16)

The uncertainty in the enthalpy at inlet and outlet are further used in the following equation to find out the uncertainty in cooling capacity.

$$\pm \,Q = \left[ {\left( {\frac{\partial Q}{{\partial \dot{m}}}\delta \dot{m}} \right)^{2} + \left( {\frac{\partial Q}{{\partial h_{\text{in}} }}\delta h_{\text{in}} } \right)^{2} + \left( {\frac{\partial Q}{{\partial h_{\text{out}} }}\delta h_{\text{out}} } \right)^{2} } \right]^{{{1 \mathord{\left/ {\vphantom {1 2}} \right. \kern-0pt} 2}}}$$
(17)

where δṁ, δhin, and δhout are the errors associated with mass flow rate and enthalpy of air at inlet and outlet, respectively. Using the average experimental data, the cooling capacity uncertainty is estimated to be

$$\pm \,Q = 0.155238^{{}}$$
$${\text{The}}\;{\text{percentage}}\;{\text{error}}\;{\text{in}}\;{\text{cooling}}\;{\text{capacity}} = \frac{ \pm \,Q}{Q} \times 100 = 4.19\%$$

*The average values of the enthalpy, humidity ratio, cooling capacity, and dehumidification factor are taken from Table 3, for uncertainty calculation.

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Nain, S., Kajal, S. & Parinam, A. Thermal performance of desiccant-based solar air-conditioning system with silica gel coating. Environ Dev Sustain 22, 281–296 (2020). https://doi.org/10.1007/s10668-018-0201-4

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Keywords

  • Dehumidification and cooling
  • Concentric tube heat exchanger
  • Adsorption heat
  • Evaporative cooler
  • Desiccant dehumidification