# Design Analysis of a Solar-Powered Water Desalination System Using Humidification Dehumidification for Continuous Water Production

^{*}

## Abstract

**:**

## 1. Introduction

^{3}h

^{−}

^{1}to 75 m

^{3}h

^{−}

^{1}. In addition, within the same air flow variation range, the performance indicator of the system, GOR, varied over ranges of 1.24 to 1.79 and 0.97 to 1.38 for the proposed system and an HDH desalination system only, respectively. Mishal [11] proposed a continuously working HDH system by feeding seawater to absorb solar energy and reserving it in a thermal storage tank with excellent insolation during the day. At night, the tank switches to discharging mode and starts feeding the HDH system with the hot seawater. The results showed that daily freshwater production is 7.6 kg per unit area of solar collector, and the daily average for GOR and the recovery ratio are 0.3 and 0.09, respectively.

^{−2}d

^{−1}for the forced convection mode, indicating the economic and feasible characteristics of the PV-driven HDH desalination system. For a CWOA HDH system with circulating air heated by solar energy, Giwa et al. [13] studied the PV panel capacity for recovery from thermal energy as well as water production, and the calculation results showed that the daily freshwater production is 2.28 Lm

^{−}

^{2}. In addition, Shiva et al. [14] proposed a novel PV/T solar humidifier and adopted it in an HDH system driven by a heat pump cycle. Under different condensation temperatures and seawater mass flow rate conditions, the thermodynamic and economic performance of the combined HDH system was investigated experimentally. The results revealed that the highest value of water evaporation rate in the humidifier was around 4.48 kg at the seawater flow rate of 0.15 kgs

^{−}

^{1}and the condensation temperature of 20 °C, and the cost of freshwater production was 0.018 USDkg

^{−}

^{1}. Elsaf [15] came up with a power and water co-generation system based on the HDH cycle and PV/T module, and the corresponding mathematical model for the integrated system was raised. The results showed that the system could produce 12 m

^{3}of freshwater and 960 kWh of electricity annually, and the unit cost for freshwater and electricity production were 0.01 USDL

^{−}

^{1}and 0.289 USDkWh

^{−}

^{1}, respectively. Therefore, compared with solar collectors, PV/T can generate electrical energy as well as heat the circulating fluid of the system, which greatly improves the overall performance of the combined HDH system.

^{−}

^{1}, respectively. Considering the recirculation of seawater and freshwater, Zarei et al. [19] raised a CACW HDH desalination system with a direct-contact humidifier and dehumidifier. The relevant theoretical and experimental analyses were conducted at the various input parameter conditions, and the results showed that the peak value of GOR and water production are 3.3 and 300 Ld

^{−}

^{1}at the equilibrium conditions of the dehumidifier.

## 2. Description of the System

- (1)
- The system operates at a steady state condition.
- (2)
- The solar radiation is uniform and stable.
- (3)
- The photoelectric conversion efficiency changes linearly with the temperature of the photovoltaic panel.
- (4)
- Heat dissipation of HDH devices is negligible.
- (5)
- The changing of kinetic and potential energy should be ignorable.
- (6)
- Fouling resistance of the HDH devices is neglected.
- (7)
- The air is always saturated along the flow path [20].
- (8)
- The fluid is in a laminar state throughout the process.

## 3. Mathematical Model of the System

#### 3.1. Model of PV/T

_{g1}+ K

_{c,g1}(T

_{c}− T

_{g1}) = K

_{g1,amb}(T

_{g1}− T

_{amb}) + K

_{g1,atm}(T

_{g1}− T

_{atm})

_{g1}refers to the absorptivity of the upper glass cover; K

_{c,g1}, K

_{g1,amb}and K

_{g1,atm}are the heat transfer coefficient between the upper glass cover and photovoltaic panel, upper glass cover and ambient space, upper glass cover and atmosphere; T

_{c}, T

_{g1}stand for the temperature of PV panel and upper glass cover; T

_{amb}and T

_{atm}represent the ambient temperature and atmospheric temperature.

_{g1}α

_{c}= K

_{c,g1}(T

_{c}− T

_{g,1}) + K

_{c,g2}(T

_{c}− T

_{g2}) + E

_{c,g2}represents the heat transfer coefficient between the lower glass cover and PV panel; E stands for the power generation through the photoelectric process, which can be provided as:

_{g1}α

_{c}η

_{c}

_{c}in Equation (3) is the photovoltaic efficiency of the PV panel, which is described as:

_{c}= η

_{ref}[1 − γ(T

_{c}− T

_{ref})]

_{ref}is the reference efficiency, T

_{ref}stands for the reference temperature according to reference efficiency, and γ denotes the attenuation coefficient. As can be obtained from Equation (4), the photovoltaic efficiency of the PV panel will decrease with the rise in its temperature, which weakens the ability to generate electricity.

_{c,g2}(T

_{c}− T

_{g2}) = K

_{g2,f}(T

_{g2}− T

_{f})

_{g2,f}refers to the heat transfer coefficient between the lower glass cover and seawater in the flow channel.

_{g2,f}A(T

_{g2}− T

_{f}) = K

_{f,amb}(A + 2e

_{f}L

_{f})(T

_{f}− T

_{amb}) + m

_{sw}C

_{p,f}(T

_{sw,2}− T

_{sw,1})

_{f,amb}stands for the heat transfer coefficient between seawater and ambient space; e

_{f}and L

_{f}represent the width and length of the fluid channel; T

_{sw,1}and T

_{sw,2}are the inlet and outlet temperature of seawater, respectively; T

_{f}denotes the mean temperature of seawater.

_{in,D}= m

_{sw}(h

_{sw,2}− h

_{sw,1})

_{sw,1}refers to the specific enthalpy of seawater at the inlet of PV/T module, h

_{sw,2}stands for the specific enthalpy of seawater at the outlet of PV/T module.

#### 3.2. Model of SC

^{−2}, T

_{sw,2}= 60 °C) are listed in Table 2.

_{h}Iα

_{g}+ K

_{f,g}(T

_{f}− T

_{g}) = K

_{g,amb}(T

_{g}− T

_{amb}) + K

_{g,atm}(T

_{g}− T

_{atm})

_{h}Iτ

_{g}α

_{c}= K

_{f,g}(T

_{f}− T

_{g}) + K

_{f,amb}(T

_{f}− T

_{amb})/S

_{h}+ m

_{sw}C

_{p,f}(T

_{Nsw,2}− T

_{Nsw,1})

_{h}denotes the concentration ratio of trough parabolic concentrator; α

_{g}refers to the absorptivity of the glass cover; K

_{f,g}, K

_{g,amb}, K

_{g,atm}, K

_{f,amb}are the heat transfer coefficient between the glass cover and seawater flow, glass cover and ambient space, glass cover and atmosphere, seawater flow and ambient space; T

_{g}stands for the temperature of glass cover; T

_{Nsw,1}and T

_{Nsw,2}are the inlet and outlet temperature of seawater, respectively; τ

_{g}represents the transmissivity of glass cover; S

_{h}represents the heat transfer area.

_{in,N}= m

_{sw}(h

_{Nsw,2}− h

_{Nsw,1})

_{Nsw,1}and h

_{Nsw,2}stand for the specific enthalpy of seawater at the inlet and outlet of SC, respectively.

#### 3.3. Model of the Recuperator

_{re}= m

_{h}(h

_{h,i}− h

_{h,o}) = m

_{c}(h

_{c,o}− h

_{c,i})

_{h}and m

_{c}represent the mass flow rate of hot and cold fluid; h

_{h,i}and h

_{c,i}refer to the specific enthalpy of hot and cold fluid at the inlet; h

_{h,o}and h

_{c,o}are the specific enthalpy of hot and cold fluid at the outlet.

_{re}= KAΔT

_{m}

_{m}refers to the log mean temperature difference between hot and cold fluid.

^{0.646}Re

^{0.583}Pr

^{1/3}

_{f,h}+ δ

_{p}/λ

_{p}+ 1/h

_{f,c}

_{f,h}and h

_{f,c}are the heat transfer coefficient of hot and cold fluid; δ

_{p}and λ

_{p}denote the thickness and thermal conductivity of the plate.

_{re}S

_{p}

_{re}represents the total plate number and S

_{p}stands for the plate area.

#### 3.4. Model of Packed Bed Humidifier and Dehumidifier

_{sw}− m

_{b}= m

_{da}(ω

_{a,2}− ω

_{a,1})

_{da}(h

_{a,2}− h

_{a,1}) = m

_{sw}h

_{sw,2}− m

_{b}h

_{b}

_{sw}, h

_{sw,2}refer to the mass flow rate and specific enthalpy of seawater at the inlet of the humidifier; m

_{b}and h

_{b}denote the mass flow rate and specific enthalpy of seawater at the outlet of the humidifier; m

_{da}is the mass flow rate of dry air; ω

_{a,1}, h

_{a,1}represent the humidity ratio and specific enthalpy of moist air inflow, and ω

_{a,2}, h

_{a,2}stand for the humidity ratio and specific enthalpy of moist air outflow.

_{pw}= m

_{fw,o}− m

_{fw,i}

_{da}(h

_{a,2}− h

_{a,1}) = m

_{fw,o}h

_{fw,2}− m

_{fw,i}h

_{fw,1}

_{fw,i}and h

_{fw,1}are mass flow rate and specific enthalpy of freshwater at the inlet of dehumidifier; m

_{fw,o}and h

_{fw,2}represent the mass flow rate and specific enthalpy of freshwater at the outlet of dehumidifier; m

_{pw}refers to water production during the condensation process.

_{max}

_{max}refers to the smaller value of the ideal maximum enthalpy change between freshwater and moist air in the dehumidifier:

_{max}= min(ΔH

_{max,fw}, ΔH

_{max,a})

_{max,fw}and ΔH

_{max,a}are the ideal maximum enthalpy change of freshwater and moist air. To intuitively express the relative size of ideal maximum enthalpy change, heat capacity ratio, HCR

_{d}, is applied to represent the ratio of them:

_{d}= ΔH

_{max,fw}/ΔH

_{max,a}

_{pw}r/Q

_{in}

## 4. Validation of the Components within the System

## 5. Result and Analysis

^{−1}for both seawater and freshwater.

#### 5.1. Parameter Analysis of PV/T Module

^{−2}, T

_{sw,2}= 60 °C, with the increase in MFRR. Apparently, the temperature of the glass cover and PV panel increases first and then decreases and reaches the maximum value when the mass flow rate ratio is equal to 2.87, MFRR = 2.87. However, the solar panel area and power generation have opposite trends that decrease first and then increase, and obtain the minimum value, E

_{min}= 481.13 kWm

^{−2}and A

_{min}= 13 m

^{2}, in the case of MFRR = 2.87. Because of the internal heat conduction process between the glass cover and the PV panel, the temperature of the glass cover changes synchronously with that of the PV panel. The lower temperature of the glass cover, on the other hand, necessitates a bigger area of PV panel as support to heat seawater to the desired temperature, which explains their opposing variation pattern. Furthermore, as the temperature rises, the photoelectric conversion efficiency of PV panels decreases, resulting in a reduction in the amount of power produced. That is why the power generation reaches its minimum value at the peak value of the PV panel temperature.

_{sw,2}= 60 °C), with the rise of MFRR, both the panel area and the total power generation decrease first and then increase, and reach their minimum value at the condition of MFRR = 2.87. Meanwhile, the enhancement of solar irradiation will increase the temperature of the PV panel, thus weakening the photoelectric conversion efficiency and, as a result, reducing the total power generation of the system. Higher incident solar radiation makes it easier for PV/T to heat seawater to a given top temperature, so the PV panel area decreases. Further, the minimum total power generation is 8.77, 7.22, and 6.25 kW, with solar irradiation increasing from 600 Wm

^{−2}to 800 and 1000 Wm

^{−2}, decreasing by 17.67 and 28.73%, and the minimum PV panel area is 28, 18, and 13 m

^{2}, with decrement rates of 35.71 and 53.57%.

#### 5.2. Parameter Analysis of HDH

^{−1}and 76.99 kJkg

^{−1}, which are both greater than the stipulated minimum enthalpy difference. Figure 5b describes the maximum ideal enthalpy change of fresh water and moist air, as well as the variation of HCR

_{d}in the dehumidifier. As previously stated, the performance of a dehumidifier is determined by the magnitude of the ideal enthalpy change of freshwater and moist air, which is expressed as the definition of HCR

_{d}. It can be found that the maximum ideal enthalpy difference of freshwater grows as the MFRR rises, but the maximum ideal enthalpy difference of moist air has the reverse pattern. Furthermore, in the case of MFRR = 2.87, where the heat capacity ratio is equal to the unit, HCR

_{d}= 1, and the maximum ideal enthalpy of freshwater is equal to that of moist air, the dehumidifier is recognized as being in the thermodynamic equilibrium.

_{d}= 1, represents the system’s optimal performance point. Moreover, for top temperatures ranging from 55 °C to 60 and 65 °C, the maximum daytime freshwater yield is 32.04, 43.1, and 53.85 kgh

^{−-1}, with growth rates of 34.5 and 68.06%; meanwhile, the highest value of water produced at night is 46.4, 57.05 and 67.47 kgh

^{−1}, increasing by 22.95 and 45.41%. Obviously, the system produces more freshwater in the night mode because the maximum water yield increases by 44.82, 32.37, and 25.29% from 55 °C to 60 and 65 °C, compared with that of the daytime. According to Figure 5b, GOR has a similar variation trend to the water productivity, reaching its peak value in the case of HCR

_{d}= 1. For top temperatures rising from 55 °C to 60 and 65 °C, the maximum value of GOR is 1.31, 1.56, and 1.71 by day, with growth rates of 19.08 and 30.53%, whereas the highest value of GOR at night is 1.044, 1.036 and 1.031, decreasing by 0.77 and 1.25%. With top temperature higher, the system in day mode tends to attain higher GOR at a larger MFRR; namely, the system performance enhances with the increase in top temperature. Despite the fact that the system generates more freshwater in the night mode at the same top temperature, the overall performance indicator, GOR, is lower than that of the daytime due to the increased energy input.

_{a,1}, decreases first, then increases, and finally stabilizes after reaching its minimum value at the balance condition point of the system. In addition, with the top temperature changing from 55 °C to 60 and 65 °C, the extremum of air inflow temperature is 41.68, 42.44 and 43.32 °C during the day, increasing by 1.82 and 3.93%, meanwhile, the lowest value of inlet air temperature at night is 32.84, 33.75 and 34.79 °C, with growth rates of 2.77 and 5.94%. As for the air outflow temperature of the humidifier, T

_{a,2}, it keeps increasing and tends to be stable with the rise of MFRR. It is obvious that the air outflow temperature at the top temperature of 60 and 65 °C in the night mode (actually, 58 and 63 °C due to 2 °C heat loss of thermal storage tank) is lower than that of air outflow at the top temperature of 55 and 60 °C in the daytime mode when the MFRR is relatively small. The reason is that the air outflow temperature of the dehumidifier running in the night mode is lower than that of daytime under the same top temperature condition; hence, the heat capacity of seawater in the humidifier is not enough in the case of a small mass flow rate ratio. Further observation of Figure 7 reveals that the mass flow ratio corresponding to the balance condition is the “turning point” where the variation of air temperature converts from a state of apparent change to a stable development stage.

_{w,1}, and freshwater outflow temperature of the dehumidifier, T

_{fw,2}. Obviously, both of them rise at first and then drop, eventually reaching the maximum value at the balance condition point. With the top temperature varying from 55 °C to 60 and 65 °C, the peak value of the seawater inflow temperature of PV/T grows from 46.94 °C to 50.9 and 54.69 °C, increasing by 8.43 and 16.51%, and the extremum of freshwater temperature at the outlet of dehumidifier rises from 51.25 °C to 54.89 and 58.34 °C, with growth rates of 7.1 and 13.83%, respectively.

#### 5.3. Energy Analysis of Integral System

_{d}is less than one, HCR

_{d}< 1, the maximum ideal enthalpy change on the air side of the dehumidifier is larger than that of the freshwater side. Hence, the temperature and humidity of the air play a leading role in the system performance, and the GOR continues to rise with the increase in MFRR. Meanwhile, when the value of HCR

_{d}is larger than one, HCR

_{d}> 1, the maximum ideal enthalpy change of the air side is lower than that of the freshwater side; thus, the performance of the system gradually degrades due to a small value of air mass flow rate, m

_{da}. At the same top temperature, the increase in air temperature and humidity during the humidification process can be fully guaranteed in any case of mass flow rate ratio. Therefore, the air outflow temperature of the humidifier, T

_{a,2}, rises continuously until the top temperature and heat exchange area limit its further development. In the dehumidifier, the freshwater yield comes from the condensation process of water vapor carried by moist air; thus the enhancement of system performance is inevitably accompanied by the drop in air outflow temperature, T

_{a,1}, and the rise in temperature of water production. When the desalination system operates at night, lower freshwater inflow temperature of the dehumidifier, T

_{fw,1}, due to a lower ambient temperature enhances the dehumidification process of moist air, hence significantly improving the water production of the system. However, the air temperature at the outlet of the dehumidifier is also less than that of the daytime mode, which burdens the humidifier as well as impairs the compensation capability of the system; thus, the MFRR corresponding to the best performance point of the system is smaller than it during the day. In addition, when the system works in daytime mode, the recuperator further correlates the variation trend of freshwater outflow temperature of dehumidifier, T

_{fw,2}, with the seawater temperature at the inlet of PV/T, T

_{w,1}. Further, the PV/T with a lower value of T

_{w,1}inevitably requires a larger PV panel area to meet the prescribed seawater outflow temperature, which causes the variation of parameters in the PV/T module.

#### 5.4. Comparison between Present System and Different HDH Systems

## 6. Conclusions

- (1)
- With a higher top temperature, the performance of the system in daytime mode enhances with a higher GOR obtained at a larger MFRR. Moreover, the system has a higher water production in night mode because the lower ambient temperature improves the dehumidification process of moist air. However, despite larger water productivity, the GOR in night mode is lower than that of the daytime due to higher energy input.
- (2)
- The MFRR corresponding to the balance condition of the system, HCR
_{d}= 1, represents the upper limit of compensation capability and the turning point of air temperature. The former denotes the increase in air temperature, and humidity can compensate for the performance degradation of the system due to a decline in air mass flow rate until the effect of too litter air flow is not negligible. For the latter, the variation of air temperature converts from a state of apparent change to a stable development stage at this point. - (3)
- Compared with other types of HDH systems, the proposed system outperforms those in prior research, which demonstrates the feasibility and superiority of the present continuous water-producing HDH system supported by thermal energy storage.
- (4)
- With different top temperatures rising from 55 °C to 60 and 65 °C, the maximum value of GOR is 1.31, 1.56, and 1.71 by day, while the highest value of GOR at night is 1.044, 1.036, and 1.031; meanwhile, the corresponding maximum daytime water productivity is 32.04, 43.1 and 53.85 kgh
^{−1}, and the highest value of water produced at night is 46.4, 57.05 and 67.47 kgh^{−1}. - (5)
- The total power generation of PV/T reaches its minimum value in the case of HCR
_{d}= 1. With solar incident irradiation increasing from 600 Wm^{−2}to 800 and 1000 Wm^{−2}, the minimal value of the total power generation is 8.77, 7.22,S and 6.25 kW, decreasing by 17.67 and 28.73%.

## Author Contributions

## Funding

## Data Availability Statement

## Conflicts of Interest

## Nomenclature

Acronyms | |

CAOW | closed-air open-water |

GOR | gained-output-ratio |

HCR | heat capacity ratio |

HDH | humidification and dehumidification |

MFRR | liquid-gas mass flow rate ratio |

PV | photovoltaic |

PV/T | photovoltaic/thermal |

SC | solar collector |

Roman symbols | |

a | specific area (m^{2}m^{−3}) |

A | heat transfer surface area (m^{2}) |

b | channel height of the plate heat exchangers (mm) |

C | concentration ratio |

C_{p} | specific heat capacity (kJkg^{−1}K^{−1}) |

e_{f} | height of flow channel (m) |

E | electric energy production (Wm^{−2}) |

h | specific enthalpy (kJkg^{−1}) |

r | latent heat (kJkg^{−1}) |

H | height (m) |

ΔH | enthalpy change (kJkg^{−1}) |

I | incident solar radiation (Wm^{−2}) |

k | mass transfer coefficient (kgm^{−2}s^{−1}) |

K | heat transfer coefficient (kWm^{−1}K^{−1}) |

L | length of photovoltaic panel (m) |

m | mass flow rate (kgs^{−1}) |

N | plate number of recuperator |

Nu | Nusselt number |

Q | heat load (kW) |

P_{r} | Prandtl number |

r | latent heat (kJkg^{−1}) |

Re | Reynolds number |

S_{h} | heat transfer area of solar collector |

ΔT_{m} | log mean temperature difference (K) |

T | temperature (°C) |

W | channel width (m) |

Greek letters | |

α | absorptivity |

β | plate chevron angle (°) |

γ | attenuation coefficient |

δ | thickness (mm) |

ε | effectiveness of dehumidifier; emissivity |

η | efficiency (%) |

λ | thermal conductivity (Wm^{−1}K^{−1}) |

τ | transmissivity |

ω | humidity ratio (kgkg^{−1}) |

Subscripts | |

a | air |

amb | ambient |

atm | atmospheric |

b | brine |

c | photovoltaic panel; cold |

d | dehumidifier |

da | dry air |

D | daytime |

E | EVA |

f | working fluid |

fw | freshwater |

g | glass cover of solar collector |

g1 | upper glass cover |

g2 | lower glass cover |

g3 | seawater flow channel |

h | hot; humidifier |

i | inlet; insulating layer |

m | main |

max | maximum |

min | minimum |

N | night |

Nsw | seawater for night use |

o | outlet |

p | plate |

pw | product water |

re | recuperator |

ref | reference |

sw | seawater |

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**Figure 1.**The schematic diagram of solar-driven HDH desalination system. (

**a**) Integral system, (

**b**) Day mode subsystem, (

**c**) Night mode subsystem.

**Figure 4.**Thermodynamic performance of PV/T module under different operation cases. (

**a**) Parameter variation under typical operations; (

**b**) Characteristics of the total power generation and PV panel area.

**Figure 5.**The energetic performance of HDH system at the prescribed conditions of I = 1000 Wm

^{−2}and T

_{sw,2}= 60 °C. (

**a**) Enthalpy versus temperature diagram of working fluid within the humidifier; (

**b**) Characteristics of ideal maximum enthalpy change and heat capacity ratio of the dehumidifier.

**Figure 6.**Schematic diagram of water production and GOR under different operation cases. (

**a**) Water production; (

**b**) GOR.

**Figure 8.**Temperature variation in seawater inflow of PV/T and freshwater outflow of dehumidifier under different operation cases.

Parameter | Data | Unit | Parameter | Data | Unit |
---|---|---|---|---|---|

$\mathrm{C}$ | 4 | - | ${\mathsf{\delta}}_{\mathrm{E}}$ | 0.0001 | m |

${\alpha}_{\mathrm{g}1}$ | 0.1 | - | ${\delta}_{\mathrm{C}}$ | 0.0003 | m |

${\tau}_{\mathrm{g}1}$ | 0.9 | - | ${\delta}_{\mathrm{i}}$ | 0.01 | m |

${\alpha}_{\mathrm{c}}$ | 0.9 | - | ${\mathsf{\lambda}}_{\mathrm{g}1}$ | 0.7 | Wm^{−1}K^{−1} |

$\gamma $ | 0.0045 | K^{−1} | ${\mathsf{\lambda}}_{\mathrm{g}2}$ | 1 | Wm^{−1}K^{−1} |

${\mathrm{T}}_{\mathrm{ref}}$ | 298.15 | K | ${\mathsf{\lambda}}_{\mathrm{g}3}$ | 0.7 | Wm^{−1}K^{−1} |

${\mathsf{\eta}}_{\mathrm{ref}}$ | 20 | % | ${\mathsf{\lambda}}_{\mathrm{E}}$ | 0.35 | Wm^{−1}K^{−1} |

${\mathsf{\delta}}_{\mathrm{g}1}$ | 0.004 | m | ${\mathsf{\lambda}}_{\mathrm{C}}$ | 148 | Wm^{−1}K^{−1} |

${\mathsf{\delta}}_{\mathrm{g}2}$ | 0.003 | m | ${\mathsf{\lambda}}_{\mathrm{i}}$ | 0.035 | Wm^{−1}K^{−1} |

${\mathsf{\delta}}_{\mathrm{g}3}$ | 0.004 | m | A | 13 | m^{2} |

e_{f} | 0.025 | m | L_{f} | 1.5 | m |

Parameter | Data | Unit | Parameter | Data | Unit |
---|---|---|---|---|---|

${\mathrm{C}}_{\mathrm{h}}$ | 10 | - | ${\mathsf{\delta}}_{\mathrm{g}}$ | 0.004 | m |

$\mathrm{I}$ | 1000 | Wm^{−2} | ${\mathsf{\delta}}_{\mathrm{i}}$ | 0.01 | m |

${\alpha}_{\mathrm{g}}$ | 0.1 | - | ${\mathsf{\lambda}}_{\mathrm{g}1}$ | 0.7 | Wm^{−1}K^{−1} |

${\tau}_{\mathrm{g}}$ | 0.9 | - | ${\mathsf{\lambda}}_{\mathrm{i}}$ | 0.035 | Wm^{−1}K^{−1} |

T_{g} | 54 | °C | S_{h} | 5.32 | m^{2} |

δ (mm) | S_{p} (mm^{2}) | W (mm) | b (mm) | β (°) |
---|---|---|---|---|

0.45 | 40,460 | 60 | 4.287 | 60 |

Term | I (Wm^{−2}) | T_{sw,1} (°C) | m_{sw} (kgs^{−1}) | E (Wm^{−2}) | η_{c} (%) |
---|---|---|---|---|---|

Calculation | 1000 | 28 | 0.028 | 407.16 | 11.54 |

Bu [21] | 1000 | 28 | 0.028 | 410.71 | 11 |

Relative error (%) | - | - | - | 0.864 | 4.91 |

Term | I (Wm^{−2}) | T_{sw,1} (°C) | m_{sw} (kgs^{−1}) | S (m^{2}) | T_{sw,2} (°C) |
---|---|---|---|---|---|

Calculation | 1000 | 106.31 | 0.028 | 2.38 | 178 |

Chen [22] | 1000 | 106.31 | 0.028 | 2.38 | 177.74 |

Relative error (%) | - | - | - | - | 0.146 |

Term | Ψ (kJkg^{−1}) | T_{a,1} (°C) | T_{sw,2} (°C) | T_{a,2} (°C) | T_{b} (°C) |
---|---|---|---|---|---|

Calculation | 20 | 25 | 80 | 67.56 | 35.89 |

Narayan [20] | 20 | 25 | 80 | 65 | 36.11 |

Relative error (%) | - | - | - | 3.94 | 0.61 |

Term | m_{sw} (kgs^{−1}) | m_{da} (kgs^{−1}) | T_{sw,2} (°C) | T_{a,2} (°C) | T_{b} (°C) |
---|---|---|---|---|---|

Calculation | 0.56 | 0.16 | 80 | 74.98 | 62.19 |

He [16] | 0.56 | 0.16 | 80 | 74.98 | 64.77 |

Relative error (%) | - | - | - | 0 | 3.98 |

Author | System | GOR | Water Productivity |
---|---|---|---|

Aref [26] | Bubble column HDH + Heat pipe | 0.81 | 8.7 Lm^{−2}day^{−1} |

Wang [27] | HDH + PCM packed bed | - | 22.5 kgh^{−1} |

Mahmoud [28] | HDH + Still | - | 11 Lm^{−2}day^{−1} |

Alrbai [29] | HDH + Fogging nozzle | 3.4 | 0.883 kgh^{−1} |

Present work | HDH + PVT + SC (Day) | 1.71 | 53.85 kgh^{−1} |

HDH + Energy stroage (Night) | 1.031 | 67.47 kgh^{−1} |

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**MDPI and ACS Style**

Lu, Y.; He, W.; Zhou, X.; An, H.; Chen, J.; Han, D.
Design Analysis of a Solar-Powered Water Desalination System Using Humidification Dehumidification for Continuous Water Production. *Sustainability* **2022**, *14*, 6860.
https://doi.org/10.3390/su14116860

**AMA Style**

Lu Y, He W, Zhou X, An H, Chen J, Han D.
Design Analysis of a Solar-Powered Water Desalination System Using Humidification Dehumidification for Continuous Water Production. *Sustainability*. 2022; 14(11):6860.
https://doi.org/10.3390/su14116860

**Chicago/Turabian Style**

Lu, Yu, Weifeng He, Xuan Zhou, Haohao An, Junjie Chen, and Dong Han.
2022. "Design Analysis of a Solar-Powered Water Desalination System Using Humidification Dehumidification for Continuous Water Production" *Sustainability* 14, no. 11: 6860.
https://doi.org/10.3390/su14116860