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Experimental investigation for a non-adiabatic desiccant wheel with a concentric structure at low regeneration temperatures
Energy Conversion and Management 201 (2019) 112165
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Experimental investigation for a non-adiabatic desiccant wheel with a concentric structure at low regeneration temperatures
T
⁎
Xingchao Zhoua, , Roger Reeceb a b
School of Photovoltaic and Renewable Energy Engineering, University of New South Wales, Sydney, NSW 2052, Australia The Commonwealth Scientific and Industrial Research Organisation (CSIRO), Energy Centre, Mayfield West, NSW 2304, Australia
A R T I C LE I N FO
A B S T R A C T
Keywords: Desiccant wheel Non-adiabatic process Dehumidification Air conditioning
The desiccant wheel-based air-conditioning system has been considered to be an effective alternative to conventional air-conditioning systems to reduce energy consumption in buildings. However, the adsorption and carryover heat released during the wheel’s working process both significantly restrict dehumidification and energy performance. Moreover, it is also difficult to apply a low temperature of thermal energy, including solar thermal energy and waste heat, in the desiccant wheel-based air-conditioning system. To improve the dehumidification and energy performance, and use a low temperature of thermal energy, this study designed, constructed and tested a non-adiabatic solid desiccant wheel, which is based on a concentric ring design. The wheel’s structure was fabricated from Nylon, using three-dimensional technology with polymer desiccant materials inserted into the channels between each of two rings. Then, cooling water was brought into the narrow passages inside the rings on the process air side of the wheel, with the aim of transforming the dehumidification process from quasi-adiabatic to quasi-isothermal. A series of experiments were then conducted to investigate this system’s performance under various operating conditions. The measurements showed that, when the inlet air and cooling water temperatures were 25 °C and 24 °C, respectively, the dehumidification performance of the new wheel was approximately the same as that of a conventional desiccant wheel. However, when the inlet temperature of the process air was 35 °C, the dehumidification process of the new wheel was very close to the ideal for a desiccant wheel – an isothermal dehumidification process. In this condition, the enthalpy effectiveness of the tested wheel was also 11% higher than that of a conventional wheel.
1. Introduction The desiccant wheel-based air-conditioning system is a promising alternative to mitigate the energy and environmental problems caused by conventional air-conditioning systems [1]. This is because the system can handle the latent cooling load via desiccant wheels driven by waste heat or solar thermal heat [2], which consumes less electricity, and can process the sensible cooling load via a dew-point evaporative cooler through the evaporation of water without using harmful refrigerants (chlorofluorocarbon refrigerants). The success of the desiccant wheel-based air-conditioning system depends largely on the thermal performance of the desiccant wheel. Conventionally, the desiccant wheel dehumidifies the air stream due to the difference in vapour pressure between the air stream and desiccant surface. This dehumidification process is approximately adiabatic, because of the release of adsorption heat in the process section of the wheel and the carryover of heat in the desiccant matrix from the regeneration section
⁎
of the wheel [3]. This adiabatic process results not only in an increase in the outlet temperature of the process air (increasing the sensible cooling load on the dew-point evaporative cooler), but also restricts the dehumidification performance of the wheel [4]. Since the 1990s, researchers have studied how desiccants shift the dehumidification processes from adiabatic to isothermal to improve the overall cooling performance – both the latent cooling capacity (i.e., better dehumidification performance) and the sensible cooling capacity (i.e., lower temperature of the air stream from the air-conditioning system) – in a non-rotating wheel-based dehumidification system. For example, Khan [5] found that heat carried over from the regeneration section of a desiccant wheel had an adverse effect on the dehumidification performance of the desiccant adsorber. Introducing a cooling source to exchange heat with desiccant could improve the dehumidification performance. Zhang et al [6] proved that, by changing the adiabatic dehumidification process to an isothermal dehumidification process, it could be possible to use a low regeneration temperature. This
Corresponding author. E-mail addresses: [email protected] (X. Zhou), [email protected] (R. Reece).
https://doi.org/10.1016/j.enconman.2019.112165 Received 12 August 2019; Received in revised form 27 September 2019; Accepted 10 October 2019 0196-8904/ © 2019 Elsevier Ltd. All rights reserved.
Energy Conversion and Management 201 (2019) 112165
X. Zhou and R. Reece
Nomenclature
h hp,out hp,out,ideal ṁ a T η
d Absolute humidity for air (g/kg dry air) D Dehumidification capacity for tested wheel (g/hour) dp,in Inlet absolute humidity for process air (g/kg dry air) dp,out Outlet absolute humidity for process air (g/kg dry air)
Enthaly for air (kj/kg dry air) Outlet enthaly for process air (kj/kg dry air) Ideal outlet enthaly for process air (kj/kg dry air) Mass flow rate for process (kg/hour) Temperature of air (°C) Enthalpy effectiveness
pump-driven two-stage desiccant wheel cooling system in Beijing summer conditions. The result for this study was that the thermal coefficient of the system’s performance was 5.5 with an inlet air humidity of 10 g/kg; cold–heat matching between the heat pump and desiccant was the key to improving performance. Although the performance of the wheel was strengthened considerably, the dehumidification process inside each wheel is still adiabatic rather than isothermal. Another issue is that the pressure drop for the system is more than double that of the pressure drop for a one-stage desiccant wheel cooling system. For an internally cooled desiccant wheel cooling system, researchers have focused on constructing a desiccant wheel with two different types of separated channels. One type of channels is used for process/regeneration air, while the other is used for bringing cooling air/water into the supply section of the wheel to take away the adsorption and regeneration heat. Studies include those of Kodama et al [13] and Narayanan et al [14], who built non-adiabatic desiccant wheels in which cooling air was used in the supply section of the desiccant wheel to eliminate the effects resulting from adsorption and carryover heat. The results of these studies showed that, compared to a conventional desiccant wheel in the same working conditions, the dehumidification performance could be improved by 30% and 45%–53% for the designs of Kodama and Narayanan, respectively. Both Goldsworthy and White [15], and Zhou et al [16] built and tested internally cooled water wheels based on a tube-shell structure, in which the cooling water was introduced into the supply section of the wheel to convert the dehumidification process from adiabatic to isothermal. The results showed that, by bringing cooling water into the wheel, the dehumidification performance could be improved significantly, by 46% and 51%, respectively, compared to a conventional wheel in the identical working conditions. However, the common characteristics for these existing non-adiabatic wheels are that the reduced face area used for processing air is quite high, at 40%[13], 50%[14], 60%[15] and 42%[16], respectively. As a result, the cooling capacity of the desiccant wheel could reduce by more than 40% when working under identical conditions; this is particularly important, because the physical size of desiccant-based airconditioners is often a key constraint when considering their application. Another constraint is that the cooling source (cooling water or air) has a cross-flow arrangement with the process air, which does not have the same heat-exchange performance as the counter-flow arrangement. The objective of this study was to design and build a more compact
could also improve the COP from 4.2 to 6.5. Yin et al [7] pointed out that an internally cooled dehumidifier could break the dehumidification limitation existed in an adiabatic one. In addition, this could also provide better dehumidification performance compared to the adiabatic one. These dehumidification processes were always achieved via an internally cooled liquid desiccant dehumidifier, which could fulfil heat exchange and dehumidification functions simultaneously. When it was processed, the air flow would have direct contact with the liquid desiccant solutions, resulting the loss of liquid desiccant solution and the deterioration of air quality. Moreover, the liquid desiccant solutions used in these studies are highly corrosive, which lead to a restrictive requirement for the material used to contain them. Thus, a solid rotating desiccant wheel could be used as an alternative to avoid these problems. In a rotating desiccant wheel design, however, approximating isothermal dehumidification in a practical device is no trivial matter. The key issues are the high pressure drop caused by the connection of multi-wheels and heat exchangers, the rapid reduction of face area used to process the air stream, the flow arrangement between the cooling source and the process air, and the difficulty of manufacturing the wheel. To solve these issues, much research work has been carried out, which can be summarised as taking two directions: one is the multi-stage desiccant wheel cooling system; and the other is the internally cooled desiccant wheel cooling system. The multi-stage desiccant cooling system mainly constitutes two desiccant wheels, two heat exchangers and two heaters, as shown in Fig. 1. The heat exchanger is located directly after the desiccant wheel and heat is exchanged between the process air and regeneration air. The adsorption and carryover heat stored in the process air can be transferred thus to the regeneration air. The dehumidification process for the desiccant wheel, combined with the heat exchanger, could therefore be considered as an isothermal process. In this process, the role of the two heaters is to heat the regeneration air to the required temperature. These studies include a two-stage desiccant wheel cooling system driven by solar energy conducted by La et al [8] and Li et al [9]. These studies found that the thermal COP could be greater than 1 and the electrical COP could be as high as 11.48; Ge et al [10,11] modelled and experimented with this two-stage desiccant wheel cooling system using different desiccant materials implemented. Their findings were that the required temperature for regeneration air was lower than the conventional desiccant wheel cooling system, which was 60 °C to 90 °C, and the cooling capacity was also 40% higher than that of a conventional desiccant cooling system. Tu et al. [12] tested the performance of a heat
Fig. 1. Schematic chart for a two-stage desiccant wheel cooling system. 2
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2.1. Description for internal heat-exchange desiccant wheel design
internally cooled desiccant wheel structure, in which cooling water could be introduced into the supply section of the wheel to change the dehumidification process from quasi-adiabatic to quasi-isothermal. In the new wheel, the face area used for processing air could be improved to about 70%. Both the dehumidification and heat-exchange functions could be fulfilled on the wheel without being influenced by the high pressure drop caused by the use of multi-wheels and heat exchangers. Moreover, a counter-flow arrangement between the cooling source and process air, providing a better heat-exchange performance than wheels with a cross-flow arrangement, could be formed due to the judicious structure. This could further strengthen the performance of the desiccant wheel. The dehumidification and energy performance of the new wheel was then implemented under various weather conditions to investigate the performance improvement. Details about the wheel design and experiment will be illustrated in Section 2, the experiment results will be demonstrated in Section 3, and the limitations of the design will be discussed in Section 4.
The design of the internal heat-exchange desiccant wheel is constituted by four identical plastic sectors (with a diameter of 0.3 m) and one aluminium central shaft. In each sector, there are 14 water channels (with a width of 5.6 mm in radius direction) used for cooling water and 13 desiccant channels (with a width of 2 mm in radius direction) between each of the two adjacent water channels used for the air streams. In this design, the total available face area used for processing air is about 0.07 m2. The water and desiccant channels are separated by a Nylon plastic wall, which was constructed using three-dimensional (3D) printing technology [17]. The thickness of the plastic wall is 1 mm. Four layers of super-adsorbent polymer desiccant material [18] were glued together and inserted into the desiccant channels. In the central shaft, there is one hole at each end side of the shaft (called the side hole) and two axial holes along the length of the shaft. The two side holes are used for the inlet and outlet of cooling water into or out of the wheel, while the two axial holes are used for the cooling water to enter or leave the process section of the wheel. Before the desiccant was combined into the system, water sealing for the surface of the plastic wall was first conducted. This was necessary because the plastic wall had a porous structure caused by the poor printing process of the 3D printing technology. Fig. 2, below, shows the water flow explanation for one sector located in the process section. In operation, the whole structure rotates around a stationary central shaft. Process air flows through the lower half of the wheel, while regeneration air flows through the upper half. These two air streams have a counter-flow arrangement. Simultaneously, cooling water enters the central shaft through one of the two side holes. It then flows into the
2. Description for the internal heat-exchange desiccant wheel design and experiment setup This section will be divided into four parts. The first part demonstrates details for the structural design of the internal heat-exchange desiccant wheel; the second part describes the test rig used to test the performance of the desiccant wheel; the third part illustrates the tested system; and the last part lists the operating conditions for the desiccant wheel during the testing process.
Fig. 2. Water and air-flow paths inside the wheel. 3
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water channels in the process section of the wheel through one of the two axial holes. After the cooling water enters the water channels, it flows continually along the water channels to the other side of the wheel in a counter-flow arrangement with the process air stream. When the cooling water reaches the end of the channel, it flows into the central shaft via the other axial hole in the shaft and leaves the wheel through the rest side hole in the shaft. Cooling water remains in the process side of the wheel due to gravity. 2.2. Description for the experimental test rig A test rig, similar to that used to test the performance of the tubeshell structure-based internally cooled desiccant wheel in the previous section, was used in the experiments for the concentric-structured based wheel, with some modifications for the cooling water loop. Again, two air streams conditioned at precisely controlled temperatures, humidity and air-flow rates from the Controlled Climate Test Facility [16] were sent through the desiccant wheel as process and regeneration air. Identical temperature and humidity sensors were used to calculate and measure heat and mass flows into and out of the tested wheel accurately. Further details on the test facility, including the detailed schematic and sensor specifications, can be found in [16]. For the cooling water loop modifications, instead of using an air–water heat exchanger to source the cooling water, a water tank integrated with a water heater and chiller was used to control the temperature of the cooling water precisely. Then, a water pump was used to adjust the flow rate of the cooling water. Fig. 3, below, shows the test
Fig. 4. Wheel used for the experiment.
loop and Fig. 4 shows the wheel used in the experiment. The reason for swapping from an air–water heat exchanger to an integrated water tank with water heater and chiller inside was that, compared with the air–water heat exchanger, an integrated water tank with water heater and chiller could control the cooling water temperature more accurately, and could also achieve a larger temperature range (< 25 °C) for cooling water.
Fig. 3. Test loop for the experiment. 4
Energy Conversion and Management 201 (2019) 112165
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Fig. 5 shows the schematic chart for the cooling water loop used in the system. The water pump sucks cooling water into the water tank for the desiccant wheel. The water flow meter is located between the wheel and water tank, and monitors the water flow. The water tank is positioned after the wheel to store the cooling water. An electrical heater and chiller connected with the water tank are used to control and maintain the water temperature to between 10 °C and 30 °C.
Table 1 Experiment test conditions for each test series. Series Process air temperature Regeneration air temperature Process air face velocity Regeneration air face velocity Rotation speed of the wheel
1
2
50 °C
25 °C 60 °C
3
4
70 °C 50 °C 2.0 m/s
5
6
35 °C 60 °C
70 °C
20 rph
2.3. Uncertainty analysis for the tested system The root sum squared method reported by Angrisani et al. [19] was adopted to analyse the measurement uncertainty. Rather than using the sum of the various uncertainty values caused by different measurements, the root sum square method evaluates the uncertainty result for a calculated experimental variable based on the root sum square of the individual uncertainties of the measured variables. The uncertainties in the measured variables would partially counteract each other most of the time, such that the square root of the sum of the squares of the individual uncertainties was a more representative value of the overall random uncertainty. This method was expressed using the following equations: 2
1/2
2
2
2 2 2⎤ ⎛ ∂f ⎞ ⎛ ∂f ⎞ ⎛ ∂f ⎞ Δy = ⎡ ⎢ ∂x1 (Δx1 ) + ∂x2 (Δx2 ) + ⋯ ∂x n (Δx n ) ⎥ ⎝ ⎠ ⎝ ⎠ ⎠ ⎝ ⎦ ⎣ ⎜
⎟
⎜
⎟
2
2
⎜
⎟
2
2
(1) 2 1/2
2
Δy ⎛ ∂f ⎞ ⎛ Δx1 ⎞ + ⎛ ∂f ⎞ ⎛ Δx2 ⎞ + ⋯⎛ ∂f ⎞ ⎛ Δx n ⎞ ⎤ =⎡ ⎥ ⎢ ∂x1 y ⎠ ⎝ y ⎠ ⎝ ∂x2 ⎠ ⎝ y ⎠ ⎝ ∂x n ⎠ ⎝ y ⎠ ⎦ ⎣⎝ ⎜
⎟
⎜
⎟
⎜
⎟
⎜
⎟
⎜
⎟
⎜
⎟
(2)
where, x i was the independent variable, Δx i was the absolute uncertainty associated with the variable x i , y was the dependent variable, Δy was its absolute uncertainty, and f was a function of the independent variable x i . For this study, the main parameters used to evaluate the performance of the internally cooled desiccant wheels were the outlet temperature and humidity ratio. They were measured by Class B resistance temperature detectors and dew-point temperature monitors, respectively. The accuracy details for these two kinds of sensors were 1.2 °C and 0.2 °C [16], respectively. Eqs. (1) and (2) were then used to calculate the relative uncertainties, which were approximately ± 12% and ± 20% for temperature and humidity, respectively.
Fig. 6. Humidity difference between inlet and outlet process air for different air face velocities and rotation speeds.
and face velocity for supply/regeneration air, and rotation speed of the wheel) are set out in Table 1. Other test conditions had the process air and regeneration air inlets’ absolute humidity set to the same measurements. The relative humidity ratios for the process air tested were 20%, 40%, 60% and 90%. The inlet temperature for cooling water was set to 15 °C to –30 °C and flow rate was set to 0–3.5 L/min (also termed the ‘dry performance’ for the new wheel whenever the water flow rate was 0 L/min). This relatively low water flow range was caused by how the wheel was constructed. Given that the four wheel sectors were 3D printed separately, and did not fit each other precisely, it was necessary to use a thin waterproof sealant (silastic) between each sector. It was then discovered that, if the water flow was too high, the high water flow could damage the waterproofing used to seal the porous structure.
2.4. Description for experimental test conditions The wheel was tested over a range of process and regeneration air inlet conditions designed to cover the typical range of operating points, as suggested by the manufacturer [20] of the conventional desiccant wheel, which covered an inlet process air temperature range of 25 °C–35 °C. Some test condition combinations (including temperature
Fig. 5. Schematic chart for the cooling loop. 5
Energy Conversion and Management 201 (2019) 112165
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Fig. 7. Thermal process change of the process air in psychrometric chart when a) inlet air temperature was 25 °C and inlet regeneration air temperature was 50 °C, b) 25 °C and 60 °C, and c) 25 °C and 70 °C. (Cooling water temperature was 24 °C and flow rate was 1.5 L/min for all three figures.)
3. Experiment results
are shown in Fig. 6. Note that, in Fig. 6, there is no recorded data for 30 rph of rotation speed and face velocity of 2.5 m/s. This was because the experimental system failed to maintain a steady state during this test. The maximum error of the outlet reading was about 1.6 g/kg (about ± 0.8 g/kg to the average value), which was far greater than the uncertainty of the typical values obtained from the experimental system. Other findings include that, when the rotation speed of the wheel was fixed (at 15, 20, 25 and 30 rph), the humidity difference between the inlet and outlet process air increased first when air face velocity varied from 1.5 to 2.0 m/s, then decreased when air face velocity increased to 3.0 m/s. The maximum difference happened in the range of 1.5–2.0 m/s air face velocity. When the air face velocity was constant, in most cases the humidity difference for 20 and 25 rph was larger than that for 15 and 30 rph. The largest humidity difference occurred when the air face velocity was 2.0 m/s for the selected wheel rotation speeds (15, 20, 25 and 30 rph). When the rotation speed was 20 rph, the average dehumidification performance was better than in the conditions with different rotation speeds. Thus, 2.0 m/s face velocity and 20 rph rotation speed were selected to do the remainder of the performance tests for the new wheel, for which the optimised rotation speed and face velocity when the internally cooled desiccant both had a relatively low process air outlet temperature and humidity (10–25 rph and 1.0–2.0 m/s).
To investigate the overall performance of the internal heat-exchange desiccant wheel, this section analyses the performance of the desiccant wheel under different conditions: face velocity of air streams, inlet temperature and humidity of air streams, and rotation speeds of desiccant wheel, for both dry and wet performance. This work will also implement a comparison between conventional desiccant wheels and internally cooled desiccant wheels to clarify the dehumidification improvement by changing from an adiabatic dehumidification process to an isothermal dehumidification process.
3.1. Analysis of results for various face velocity of air streams and rotation speeds of the concentric design wheel To ensure that the concentric design wheel could work in the optimal air face velocity and rotation speed range, the optimal air face velocity and rotation speed of the wheel was determined before the series of tests began. Here, the optimal air face velocity and rotation speed were those in which the desiccant wheel had the maximum humidity difference between the process air inlet and outlet. The test conditions for these investigations were as follows: process and regeneration air temperatures were between 30 °C and 50 °C, respectively; air face velocity varied from 1.5 to 3.0 m/s (process and regeneration air had the same face velocity); wheel’s rotation speed varied from 15 to 30 rph (resolution per hour); temperature for cooling water was 24 °C; and flow rate was 1.5 L/min. The experiment results 6
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Fig. 8. Thermal process change of the process air in psychrometric chart when a) inlet air temperature was 35 °C and inlet regeneration air temperature is 50 °C, b) 35 °C and 60 °C, and c) 35 °C and 70 °C. (Cooling water temperature was 24 °C and flow rate was 1.5 L/min for all three figures.)
(for wet performance) than the outlet temperature for a conventional desiccant wheel. The small temperature difference between the outlet temperatures for the wet performance of the new wheel and the performance for a conventional desiccant wheel was caused by the tiny temperature difference of the cooling water (24 °C) and process air inlet’s temperature (25 °C). For the process air outlet’s humidity, the results for dry performance was, on average, 1.1 g/kg higher than that of conventional performance, while the results for wet performance was, on average, 0.6 g/kg higher than that of conventional performance. The process air outlet’s humidity in wet performance fluctuated around the results of conventional performance. This meant that the small temperature difference between the cooling water and process air inlet did not influence the dehumidification performance of the wheel too much when the process air inlet’s temperature was 25 °C. However, for the process air’s outlet temperature, when the process air inlet’s temperature was 35 °C, the process air outlet’s temperature in dry performance was also slightly higher (around 0.6 in average) than that of conventional performance while the process air temperature in wet performance was, on average, 7 °C lower than that of conventional performance. This meant that, when the temperature difference between the cooling water (24 °C) and the process air inlet (35 °C) became larger, the thermal performance (regarding the process air outlet’s temperature) became better. The process air outlet’s humidity, for dry performance, was 0.4 g/kg higher than that for conventional performance while the outlet humidity for wet performance was, on average, 0.4 g/kg lower than that
3.2. Analysis of results for various inlet temperature and humidity of air streams Fig. 7(a), (b) and (c) show the changing thermal conditions of the process air in a psychrometric chart, when the process and regeneration air inlet temperatures were 25 °C/50 °C, 25 °C/60 °C and 25 °C/70 °C, respectively. Fig. 8(a), (b) and (c) show the changing thermal conditions of the process air in a psychrometric chart, when the process air/ regeneration air inlet temperatures were 35 °C/50 °C, 35 °C/60 °C and 35 °C/70 °C for conventional desiccant wheels (green rotor wheels with a quasi-adiabatic adiabatic dehumidification process using results data provided by the manufacturer [20]) in both the dry performance for the new wheel (with the test results for the new wheel when it worked while cooling water did not flow inside it) and the wet performance for the concentric structure-based wheel (with the test results when it worked while cooling water flowed inside it). The desiccant materials used in these wheels were exactly same, including their isothermal properties and channel shapes. The results in Figs. 7 and 8 show that, when the inlet temperature of process air was 25 °C, the process air outlet’s temperature in conventional performance was, on average, 2.4 °C lower than that of the dry performance, while the process air outlet’s temperature in the wet performance was, on average, 3 °C lower than that of conventional performance. This implies two things: the first was that the extra material used to construct the new wheel did not negatively impact the performance of the wheel too much due to its thermal capacity. The second was that the new wheel could have a lower outlet temperature 7
Energy Conversion and Management 201 (2019) 112165
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Fig. 9. Dehumidification performance for dry and wet performance of the concentric design wheel when inlet process air temperature was a) 25 °C and b) 35 °C.
Angrisani et al. [19] in 2012. ii. The dehumidification performance for the wet process was always better than for the dry process (even in the same inlet conditions), and the average increase was about 0.11 kg/h.
of conventional performance. However, the theoretical humidity difference between the process air inlet and outlet from the simulated results after optimisation in [16] was more than 1.4 g/kg. Thus, the improvement of dehumidification performance (in wet performance) for the new wheel was far from ideal. The potential reasons for this will be explained in the discussion section.
3.4. Performance improvement analysis for the concentric design wheel based on the enthalpy effectiveness
3.3. Analysis of results for dehumidification performance of concentric design wheel in dry and wet performance
Both the process air outlet’s temperature and humidity need to be considered when evaluating the performance improvement of the desiccant wheel, changing from an adiabatic dehumidification process to an isothermal dehumidification process. Thus, a new evaluation standard proposed by Mandageri [21] was adopted to evaluate the overall performance improvement of the concentric design wheel compared to a conventional desiccant wheel. This new standard was termed the ‘enthalpy effectiveness’ of the desiccant wheel and showed the deviation of the outlet process air enthalpy from the inlet air enthalpy, which accounted for both outlet temperature and humidity (defined below):
The dehumidification performance of the desiccant wheel was defined using the following equation:
D = ṁ a (dp, in − dp, out )
(3)
Fig. 9(a) and (b) show the dehumidification performance of the new wheel (both dry and wet performance) when the cooling water temperature was 24 °C, flow rate was 1.5 L/min, wheel’s rotation speed was 20 rph, air face velocity for both process air and regeneration air was 2.0 m/s, and the inlet absolute humidity ratio for both process and regeneration air was the same. Two main findings can be concluded from Fig. 9(a) and (b): i. When the temperature of both the process and regeneration air was kept constant, the concentric design-based wheel’s dehumidification performance increased as the inlet air relative humidity ratio increased. This occurred under both dry conditions (no water cooling) and wet conditions (with water cooling). When the inlet temperature of the process air was kept constant, the concentric design-based wheel’s dehumidification performance improved as the regeneration air temperature increased. This finding is in agreement with that of a conventional desiccant wheel, as reported by
η=1−
hp, out − hp, out , ideal hp, out , ideal
h = 1.01T + (2500 + 1.84T ) d
(4) (5)
Eq. (4) shows that, when the process air outlet’s enthalpy was the ideal value (this was the value when the process air outlet worked in an isothermal process, which can be calculated using the model introduced in [16]), the enthalpy effectiveness is 1. The closer the process air outlet’s enthalpy is to the ideal enthalpy, the larger enthalpy effectiveness it will have. However, with the conventional desiccant wheel, the released heat due to the adsorption of water vapour by the 8
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Fig. 10. Enthalpy effectiveness for the concentric design wheel when temperature for process/regeneration air was: a) 25/50 °C, b) 25/60 °C and c) 25/70 °C (Cooling water temperature was 24 °C and flow rate was 1.5 L/min for all three figures).
process air inlet’s humidity ratio increased.
desiccant, and the heat carried over from the regeneration section and stored in the matrix, would lead to a quasi-adiabatic dehumidification process, which provides an effectiveness of 0.35–0.84 for the superadsorbent desiccant used in this study (the effectiveness is calculated out based on the data provided by [16,20]). Figs. 10 and 11 show the enthalpy effectiveness of the concentric design wheel under varying weather conditions, for both dry and wet performance, along with the enthalpy effectiveness of the conventional desiccant wheel (as mentioned above). As shown in Fig. 10, the conventional desiccant wheel’s enthalpy effectiveness was higher than that of the new wheel when it worked in dry performance mode. This implies that the structural material used to form the cooling water and desiccant channel could create extra thermal mass due to its thermal capacity. This extra thermal mass had a negative influence on the performance of the wheel. The enthalpy effectiveness for the new wheel when it worked in wet performance mode was always slightly higher than that of the conventional wheel. This implies that the introduction of cooling water into the wheel’s process section could strengthen the wheel’s enthalpy effectiveness. However, this improvement was not obvious (an average increase of 4%), due to the small temperature difference between the cooling water and the process air. The finding concluded from Fig. 11 was that, when the process air inlet’s temperature was 35 °C, the wheel’s enthalpy effectiveness in wet performance mode was much higher than that of the conventional wheel (an average increase of 11%), but it was still lower than that of the conventional wheel when it worked in dry performance mode. When the process air inlet’s temperature was at 25 °C and 35 °C, the enthalpy effectiveness decreased as the regeneration temperature and
3.5. Performance comparisons between conventional desiccant wheels and internally cooled desiccant wheels To better understand the performance of the conventional desiccant wheel (using different materials, including zeolite, silica gel and polymer desiccant) and internally cooled desiccant wheels (with tubeshell and concentric structures) regenerated by a low regeneration temperature (50 °C), the performance result comparisons are given in Fig. 12. The setting conditions were as follows: rotation speed of all desiccant wheels was 20 rph; inlet temperature for the cooling water was 24 °C, and flow rate was 4.5 L/min for the internally cooled desiccant wheels; inlet temperature was 30 °C for the supply air and 50 °C for the regeneration air, for all desiccant wheels; and relative humidity ratio varied from 30% to 90% for the supply air flow (the regeneration air had the same absolute humidity as the supply air), for all desiccant wheels. Because conventional and internally cooled desiccant wheels both have different optimal face velocities, a face velocity of 2.5 m/s [22] was set for both supply and regeneration air flow for the conventional desiccant wheels, while a face velocity of 2.0 m/s [15,16,23] was set for both supply and regeneration air flow of the internally cooled desiccant wheels. As can be seen in Fig. 12, when using low temperature of regenerative heat, the performance of the conventional desiccant wheel using a zeolite desiccant was not as good as the other wheels, because it had the highest outlet temperature and humidity. The outlet temperature of the conventional desiccant wheel using silica gel desiccant was always higher than that using a polymer desiccant. While the outlet 9
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Fig. 11. Enthalpy effectiveness for the concentric wheel when temperature for process/regeneration air was: a) 35/50 °C, b) 35/60 °C and c) 35/70 °C (Cooling water temperature was 24 °C and flow rate was 1.5 L/min for all three figures).
0.4 g/kg) was not as good as the simulated dehumidification improvement (on average, 0.79 g/kg). There were three reasons that can explain this unexpected performance. First, the width of the channels used for cooling water was too narrow (the channels were less than 2 mm, which was narrower than the channels in the tube-shell design internally cooled desiccant wheel), due to an inaccuracy in the 3D printing technology. When the desiccant channels in the wheel’s process section rotated to the regeneration section, the cooling water did not have enough time to fall into the process section, which could then be brought to the wheel’s regeneration section and exchange heat with the regeneration air. The second reason was that, due to the fragile waterproof sealant, the flow rate of the cooling water was too low (the maximum water flow rate was only 3.5 L/min); this significantly influenced the heatexchange process between the process air and cooling water. The final reason was that the desiccant wheel had a porous structure due to the poor 3D printing technology, which led to water leakage. Although surface sealing work had been performed before the formal test, there was still a small amount of water leakage (which could increase the air humidity ratio about 1.2 g/kg) from water channels into the desiccant channels, which could weaken the new wheel’s dehumidification performance. To improve the performance of the internally cooled desiccant wheel, a more accurate manufacturing process is required to produce the concentric structure. In addition, the use of nonporous materials and a simpler wheel structure would also be advantageous. These approaches are recommended as future work, which would need to be done to move the technology towards a commercial product.
humidity of the conventional desiccant wheel using silica gel desiccant was higher than that using polymer desiccant, this was only when the humidity was high (i.e., relative humidity ratio greater than 60% [24]). The performance of the internally cooled desiccant wheel with a tubeshell structure was only a small improvement to the conventional desiccant wheel using polymer desiccant, due to the high heat resistance caused by the overlap of multiple desiccant layers. However, the performance of the internally cooled desiccant wheel with a concentric structure improved significantly compared to the conventional desiccant wheel using a polymer desiccant, because the outlet temperature and humidity was more than 3 °C lower and 1 g/kg higher than those of the conventional desiccant wheel using a polymer desiccant. 4. Discussion This study investigated the cooling and dehumidification performance of a new internally cooled desiccant wheel with a counter-flow configuration between the cooling water and process air. In the new wheel, the high heat resistance caused by the overlap of too many desiccant layers was avoided. This was because, instead of wrapping many desiccant layers in a column inside a heat-exchange tube, the desiccant layers were attached inside concentric rings around the water channels. The maximum number of desiccant layers between any two water channels was only four. The ratio of the face area used for processing air was about 67% of the total face area. Another advantage of this wheel compared to existing non-adiabatic wheels was that the new wheel had a counter-flow arrangement between the process and the cooling source, which had a better heat-exchange performance than the crossflow arrangements in existing non-adiabatic wheels. However, the overall dehumidification improvement (on average, 10
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Fig. 12. Outlet temperature (a) and humidity (b) comparisons between conventional desiccant wheels (using zeolite, silica gel and polymer desiccant [22]) and internally cooled desiccant wheels (based on tube-shell and concentric designs).
5. Conclusion
The new wheel design uses the lowest temperature (50 °C–70 °C) of thermal energy possible, including solar thermal and waste heat. Thus, the next stage of work will be to analyse the feasibility for the application of the desiccant wheel-based cooling system using the internally cooled wheel in various climate areas.
In this study, an internally cooled desiccant wheel-based on a concentric structure was designed, constructed, tested and analysed. Compared to the tube-shell design [16], this design did not have multiple desiccant layers overlapping each other, so the high heat resistance could be avoided. The water flow in this design had a counterflow arrangement with the process air; this had a better heat-exchange performance than that present in the tube-shell design. The new wheel’s performance was tested for various weather conditions; the results were plotted in a psychrometric chart and compared to a conventional desiccant wheel’s performance to evaluate its dehumidification process. During the evaluation, enthalpy effectiveness was applied to analyse the non-adiabatic performance of both wheels. The main findings can be summarised as follows:
Declaration of Competing Interest The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper. Acknowledgements This project is part of the desiccant cooling program conducted by the University of New South Wales and CSIRO and funded by the CRC for Low Carbon Living Ltd, which is supported by the Cooperative Research Centres program, an Australian Government initiative.
1. When the temperature difference between the process air (25 °C) and cooling water (24 °C) was too small, the thermal performance of the concentric design wheel did not improve significantly. When this difference became larger (24 °C for cooling water and 35 °C for process air), the thermal process of the new wheel improved significantly, shifting from an adiabatic process to a semi-isothermal process. 2. The concentric design wheel’s enthalpy effectiveness decreased with an increase in the process air inlet’s relative humidity and the regeneration air temperature. 3. The concentric design wheel’s enthalpy effectiveness was higher than that of the conventional desiccant wheel (on average, 11% and 4% higher for process air inlet temperatures of 35 °C and 25 °C, respectively) when it worked in wet performance mode. It was lower (on average, 3% and 13% lower for process air inlet temperatures of 35 °C and 25 °C, respectively) than that of the conventional desiccant wheel when it worked in dry performance mode.
References [1] Guo J, Lin S, Bilbao JI, White SD, Sproul AB. A review of photovoltaic thermal (PV/ T) heat utilisation with low temperature desiccant cooling and dehumidification. Renew Sustain Energy Rev 2017;67:1–14. [2] Jeong J, Yamaguchi S, Saito K, Kawai S. Performance analysis of four-partition desiccant wheel and hybrid dehumidification air-conditioning system. Int J Refrig. 2010;33:496–509. [3] La D, Dai YJ, Li Y, Wang RZ, Ge TS. Technical development of rotary desiccant dehumidification and air conditioning: a review. Renew Sustain Energy Rev 2010;14:130–47. [4] Huan Z, Jianlei N. A two-stage desiccant cooling system using low-temperature heat. Building Serv Eng Res Technol. 1999;20:51–5. [5] Khan AY. Cooling and dehumidification performance analysis of internally-cooled liquid desiccant absorbers. Appl Therm Eng. 1998;18:265–81. [6] Zhang T, Liu X, Jiang J, Chang X, Jiang Y. Experimental analysis of an internallycooled liquid desiccant dehumidifier. BBuild Environ. 2013;63:1–10. [7] Yin Y, Zhang X, Peng D, Li X. Model validation and case study on internally cooled/
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[15] Goldsworthy MJ, White S. Design and performance of an internal heat exchange desiccant wheel. Int J Refrig 2014;39:152–9. [16] Zhou X, Goldsworthy M, Sproul A. Performance investigation of an internally cooled desiccant wheel. Appl Energy 2018;224:382–97. [17] Printers ESD, http://3dprintingsystems.com/3d_printers/shining3dprinters/ [accessed 20 November 2018]. [18] Lee J, Lee D-Y. Sorption characteristics of a novel polymeric desiccant. Int J Refrig 2012;35:1940–9. [19] Angrisani G, Minichiello F, Roselli C, Sasso M. Experimental analysis on the dehumidification and thermal performance of a desiccant wheel. Appl Energy 2012;92:563–72. [20] SDP Desiccant Rotor, http://greenrotors.com/ [accessed 20 November 2018]. [21] Ali Mandegari M, Pahlavanzadeh H. Introduction of a new definition for effectiveness of desiccant wheels. Energy 2009;34:797–803. [22] White SD, Goldsworthy M, Reece R, Spillmann T, Gorur A, Lee D-Y. Characterization of desiccant wheels with alternative materials at low regeneration temperatures. Int J Refrig. 2011;34:1786–91. [23] Zhou X, Goldsworthy M, Sproul A. Application of internal cooled desiccant wheel in air-conditioning system. ASHRAE J 2018. [24] Ahlberg JE. Rates of Water Vapor Adsorption from Air by Silica Gel. Ind Eng Chem Res. 1939;31:988–92.
heated dehumidifier/regenerator of liquid desiccant systems. Int J Therm Sci. 2009;48:1664–71. La D, Dai Y, Li Y, Ge T, Wang R. Case study and theoretical analysis of a solar driven two-stage rotary desiccant cooling system assisted by vapor compression air-conditioning. J Sol. 2011;85:2997–3009. Li H, Dai YJ, Li Y, La D, Wang RZ. Case study of a two-stage rotary desiccant cooling/heating system driven by evacuated glass tube solar air collectors. Energy Build. 2012;47:107–12. Ge TS, Li Y, Wang RZ, Dai YJ. Experimental study on a two-stage rotary desiccant cooling system. Int J Refrig. 2009;32:498–508. Ge TS, Li Y, Dai YJ, Wang RZ. Performance investigation on a novel two-stage solar driven rotary desiccant cooling system using composite desiccant materials. J Sol. 2010;84:157–9. Tu R, Liu X-H, Hwang Y, Ma F. Performance analysis of ventilation systems with desiccant wheel cooling based on exergy destruction. Energy Convers Manage 2016;123:265–79. Kodama A, Watanabe N, Hirose T, Goto M, Okano H. Performance of a multipass honeycomb adsorber regenerated by a direct hot water heating. Adsorption. 2005;11:603–8. Narayanan R, Saman WY, White SD. A non-adiabatic desiccant wheel: Modeling and experimental validation. Appl Therm Eng 2013;61:178–85.
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Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Experimental investigation for a non-adiabatic desiccant wheel with a concentric structure at low regeneration temperatures
T
⁎
Xingchao Zhoua, , Roger Reeceb a b
School of Photovoltaic and Renewable Energy Engineering, University of New South Wales, Sydney, NSW 2052, Australia The Commonwealth Scientific and Industrial Research Organisation (CSIRO), Energy Centre, Mayfield West, NSW 2304, Australia
A R T I C LE I N FO
A B S T R A C T
Keywords: Desiccant wheel Non-adiabatic process Dehumidification Air conditioning
The desiccant wheel-based air-conditioning system has been considered to be an effective alternative to conventional air-conditioning systems to reduce energy consumption in buildings. However, the adsorption and carryover heat released during the wheel’s working process both significantly restrict dehumidification and energy performance. Moreover, it is also difficult to apply a low temperature of thermal energy, including solar thermal energy and waste heat, in the desiccant wheel-based air-conditioning system. To improve the dehumidification and energy performance, and use a low temperature of thermal energy, this study designed, constructed and tested a non-adiabatic solid desiccant wheel, which is based on a concentric ring design. The wheel’s structure was fabricated from Nylon, using three-dimensional technology with polymer desiccant materials inserted into the channels between each of two rings. Then, cooling water was brought into the narrow passages inside the rings on the process air side of the wheel, with the aim of transforming the dehumidification process from quasi-adiabatic to quasi-isothermal. A series of experiments were then conducted to investigate this system’s performance under various operating conditions. The measurements showed that, when the inlet air and cooling water temperatures were 25 °C and 24 °C, respectively, the dehumidification performance of the new wheel was approximately the same as that of a conventional desiccant wheel. However, when the inlet temperature of the process air was 35 °C, the dehumidification process of the new wheel was very close to the ideal for a desiccant wheel – an isothermal dehumidification process. In this condition, the enthalpy effectiveness of the tested wheel was also 11% higher than that of a conventional wheel.
1. Introduction The desiccant wheel-based air-conditioning system is a promising alternative to mitigate the energy and environmental problems caused by conventional air-conditioning systems [1]. This is because the system can handle the latent cooling load via desiccant wheels driven by waste heat or solar thermal heat [2], which consumes less electricity, and can process the sensible cooling load via a dew-point evaporative cooler through the evaporation of water without using harmful refrigerants (chlorofluorocarbon refrigerants). The success of the desiccant wheel-based air-conditioning system depends largely on the thermal performance of the desiccant wheel. Conventionally, the desiccant wheel dehumidifies the air stream due to the difference in vapour pressure between the air stream and desiccant surface. This dehumidification process is approximately adiabatic, because of the release of adsorption heat in the process section of the wheel and the carryover of heat in the desiccant matrix from the regeneration section
⁎
of the wheel [3]. This adiabatic process results not only in an increase in the outlet temperature of the process air (increasing the sensible cooling load on the dew-point evaporative cooler), but also restricts the dehumidification performance of the wheel [4]. Since the 1990s, researchers have studied how desiccants shift the dehumidification processes from adiabatic to isothermal to improve the overall cooling performance – both the latent cooling capacity (i.e., better dehumidification performance) and the sensible cooling capacity (i.e., lower temperature of the air stream from the air-conditioning system) – in a non-rotating wheel-based dehumidification system. For example, Khan [5] found that heat carried over from the regeneration section of a desiccant wheel had an adverse effect on the dehumidification performance of the desiccant adsorber. Introducing a cooling source to exchange heat with desiccant could improve the dehumidification performance. Zhang et al [6] proved that, by changing the adiabatic dehumidification process to an isothermal dehumidification process, it could be possible to use a low regeneration temperature. This
Corresponding author. E-mail addresses: [email protected] (X. Zhou), [email protected] (R. Reece).
https://doi.org/10.1016/j.enconman.2019.112165 Received 12 August 2019; Received in revised form 27 September 2019; Accepted 10 October 2019 0196-8904/ © 2019 Elsevier Ltd. All rights reserved.
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Nomenclature
h hp,out hp,out,ideal ṁ a T η
d Absolute humidity for air (g/kg dry air) D Dehumidification capacity for tested wheel (g/hour) dp,in Inlet absolute humidity for process air (g/kg dry air) dp,out Outlet absolute humidity for process air (g/kg dry air)
Enthaly for air (kj/kg dry air) Outlet enthaly for process air (kj/kg dry air) Ideal outlet enthaly for process air (kj/kg dry air) Mass flow rate for process (kg/hour) Temperature of air (°C) Enthalpy effectiveness
pump-driven two-stage desiccant wheel cooling system in Beijing summer conditions. The result for this study was that the thermal coefficient of the system’s performance was 5.5 with an inlet air humidity of 10 g/kg; cold–heat matching between the heat pump and desiccant was the key to improving performance. Although the performance of the wheel was strengthened considerably, the dehumidification process inside each wheel is still adiabatic rather than isothermal. Another issue is that the pressure drop for the system is more than double that of the pressure drop for a one-stage desiccant wheel cooling system. For an internally cooled desiccant wheel cooling system, researchers have focused on constructing a desiccant wheel with two different types of separated channels. One type of channels is used for process/regeneration air, while the other is used for bringing cooling air/water into the supply section of the wheel to take away the adsorption and regeneration heat. Studies include those of Kodama et al [13] and Narayanan et al [14], who built non-adiabatic desiccant wheels in which cooling air was used in the supply section of the desiccant wheel to eliminate the effects resulting from adsorption and carryover heat. The results of these studies showed that, compared to a conventional desiccant wheel in the same working conditions, the dehumidification performance could be improved by 30% and 45%–53% for the designs of Kodama and Narayanan, respectively. Both Goldsworthy and White [15], and Zhou et al [16] built and tested internally cooled water wheels based on a tube-shell structure, in which the cooling water was introduced into the supply section of the wheel to convert the dehumidification process from adiabatic to isothermal. The results showed that, by bringing cooling water into the wheel, the dehumidification performance could be improved significantly, by 46% and 51%, respectively, compared to a conventional wheel in the identical working conditions. However, the common characteristics for these existing non-adiabatic wheels are that the reduced face area used for processing air is quite high, at 40%[13], 50%[14], 60%[15] and 42%[16], respectively. As a result, the cooling capacity of the desiccant wheel could reduce by more than 40% when working under identical conditions; this is particularly important, because the physical size of desiccant-based airconditioners is often a key constraint when considering their application. Another constraint is that the cooling source (cooling water or air) has a cross-flow arrangement with the process air, which does not have the same heat-exchange performance as the counter-flow arrangement. The objective of this study was to design and build a more compact
could also improve the COP from 4.2 to 6.5. Yin et al [7] pointed out that an internally cooled dehumidifier could break the dehumidification limitation existed in an adiabatic one. In addition, this could also provide better dehumidification performance compared to the adiabatic one. These dehumidification processes were always achieved via an internally cooled liquid desiccant dehumidifier, which could fulfil heat exchange and dehumidification functions simultaneously. When it was processed, the air flow would have direct contact with the liquid desiccant solutions, resulting the loss of liquid desiccant solution and the deterioration of air quality. Moreover, the liquid desiccant solutions used in these studies are highly corrosive, which lead to a restrictive requirement for the material used to contain them. Thus, a solid rotating desiccant wheel could be used as an alternative to avoid these problems. In a rotating desiccant wheel design, however, approximating isothermal dehumidification in a practical device is no trivial matter. The key issues are the high pressure drop caused by the connection of multi-wheels and heat exchangers, the rapid reduction of face area used to process the air stream, the flow arrangement between the cooling source and the process air, and the difficulty of manufacturing the wheel. To solve these issues, much research work has been carried out, which can be summarised as taking two directions: one is the multi-stage desiccant wheel cooling system; and the other is the internally cooled desiccant wheel cooling system. The multi-stage desiccant cooling system mainly constitutes two desiccant wheels, two heat exchangers and two heaters, as shown in Fig. 1. The heat exchanger is located directly after the desiccant wheel and heat is exchanged between the process air and regeneration air. The adsorption and carryover heat stored in the process air can be transferred thus to the regeneration air. The dehumidification process for the desiccant wheel, combined with the heat exchanger, could therefore be considered as an isothermal process. In this process, the role of the two heaters is to heat the regeneration air to the required temperature. These studies include a two-stage desiccant wheel cooling system driven by solar energy conducted by La et al [8] and Li et al [9]. These studies found that the thermal COP could be greater than 1 and the electrical COP could be as high as 11.48; Ge et al [10,11] modelled and experimented with this two-stage desiccant wheel cooling system using different desiccant materials implemented. Their findings were that the required temperature for regeneration air was lower than the conventional desiccant wheel cooling system, which was 60 °C to 90 °C, and the cooling capacity was also 40% higher than that of a conventional desiccant cooling system. Tu et al. [12] tested the performance of a heat
Fig. 1. Schematic chart for a two-stage desiccant wheel cooling system. 2
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2.1. Description for internal heat-exchange desiccant wheel design
internally cooled desiccant wheel structure, in which cooling water could be introduced into the supply section of the wheel to change the dehumidification process from quasi-adiabatic to quasi-isothermal. In the new wheel, the face area used for processing air could be improved to about 70%. Both the dehumidification and heat-exchange functions could be fulfilled on the wheel without being influenced by the high pressure drop caused by the use of multi-wheels and heat exchangers. Moreover, a counter-flow arrangement between the cooling source and process air, providing a better heat-exchange performance than wheels with a cross-flow arrangement, could be formed due to the judicious structure. This could further strengthen the performance of the desiccant wheel. The dehumidification and energy performance of the new wheel was then implemented under various weather conditions to investigate the performance improvement. Details about the wheel design and experiment will be illustrated in Section 2, the experiment results will be demonstrated in Section 3, and the limitations of the design will be discussed in Section 4.
The design of the internal heat-exchange desiccant wheel is constituted by four identical plastic sectors (with a diameter of 0.3 m) and one aluminium central shaft. In each sector, there are 14 water channels (with a width of 5.6 mm in radius direction) used for cooling water and 13 desiccant channels (with a width of 2 mm in radius direction) between each of the two adjacent water channels used for the air streams. In this design, the total available face area used for processing air is about 0.07 m2. The water and desiccant channels are separated by a Nylon plastic wall, which was constructed using three-dimensional (3D) printing technology [17]. The thickness of the plastic wall is 1 mm. Four layers of super-adsorbent polymer desiccant material [18] were glued together and inserted into the desiccant channels. In the central shaft, there is one hole at each end side of the shaft (called the side hole) and two axial holes along the length of the shaft. The two side holes are used for the inlet and outlet of cooling water into or out of the wheel, while the two axial holes are used for the cooling water to enter or leave the process section of the wheel. Before the desiccant was combined into the system, water sealing for the surface of the plastic wall was first conducted. This was necessary because the plastic wall had a porous structure caused by the poor printing process of the 3D printing technology. Fig. 2, below, shows the water flow explanation for one sector located in the process section. In operation, the whole structure rotates around a stationary central shaft. Process air flows through the lower half of the wheel, while regeneration air flows through the upper half. These two air streams have a counter-flow arrangement. Simultaneously, cooling water enters the central shaft through one of the two side holes. It then flows into the
2. Description for the internal heat-exchange desiccant wheel design and experiment setup This section will be divided into four parts. The first part demonstrates details for the structural design of the internal heat-exchange desiccant wheel; the second part describes the test rig used to test the performance of the desiccant wheel; the third part illustrates the tested system; and the last part lists the operating conditions for the desiccant wheel during the testing process.
Fig. 2. Water and air-flow paths inside the wheel. 3
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water channels in the process section of the wheel through one of the two axial holes. After the cooling water enters the water channels, it flows continually along the water channels to the other side of the wheel in a counter-flow arrangement with the process air stream. When the cooling water reaches the end of the channel, it flows into the central shaft via the other axial hole in the shaft and leaves the wheel through the rest side hole in the shaft. Cooling water remains in the process side of the wheel due to gravity. 2.2. Description for the experimental test rig A test rig, similar to that used to test the performance of the tubeshell structure-based internally cooled desiccant wheel in the previous section, was used in the experiments for the concentric-structured based wheel, with some modifications for the cooling water loop. Again, two air streams conditioned at precisely controlled temperatures, humidity and air-flow rates from the Controlled Climate Test Facility [16] were sent through the desiccant wheel as process and regeneration air. Identical temperature and humidity sensors were used to calculate and measure heat and mass flows into and out of the tested wheel accurately. Further details on the test facility, including the detailed schematic and sensor specifications, can be found in [16]. For the cooling water loop modifications, instead of using an air–water heat exchanger to source the cooling water, a water tank integrated with a water heater and chiller was used to control the temperature of the cooling water precisely. Then, a water pump was used to adjust the flow rate of the cooling water. Fig. 3, below, shows the test
Fig. 4. Wheel used for the experiment.
loop and Fig. 4 shows the wheel used in the experiment. The reason for swapping from an air–water heat exchanger to an integrated water tank with water heater and chiller inside was that, compared with the air–water heat exchanger, an integrated water tank with water heater and chiller could control the cooling water temperature more accurately, and could also achieve a larger temperature range (< 25 °C) for cooling water.
Fig. 3. Test loop for the experiment. 4
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Fig. 5 shows the schematic chart for the cooling water loop used in the system. The water pump sucks cooling water into the water tank for the desiccant wheel. The water flow meter is located between the wheel and water tank, and monitors the water flow. The water tank is positioned after the wheel to store the cooling water. An electrical heater and chiller connected with the water tank are used to control and maintain the water temperature to between 10 °C and 30 °C.
Table 1 Experiment test conditions for each test series. Series Process air temperature Regeneration air temperature Process air face velocity Regeneration air face velocity Rotation speed of the wheel
1
2
50 °C
25 °C 60 °C
3
4
70 °C 50 °C 2.0 m/s
5
6
35 °C 60 °C
70 °C
20 rph
2.3. Uncertainty analysis for the tested system The root sum squared method reported by Angrisani et al. [19] was adopted to analyse the measurement uncertainty. Rather than using the sum of the various uncertainty values caused by different measurements, the root sum square method evaluates the uncertainty result for a calculated experimental variable based on the root sum square of the individual uncertainties of the measured variables. The uncertainties in the measured variables would partially counteract each other most of the time, such that the square root of the sum of the squares of the individual uncertainties was a more representative value of the overall random uncertainty. This method was expressed using the following equations: 2
1/2
2
2
2 2 2⎤ ⎛ ∂f ⎞ ⎛ ∂f ⎞ ⎛ ∂f ⎞ Δy = ⎡ ⎢ ∂x1 (Δx1 ) + ∂x2 (Δx2 ) + ⋯ ∂x n (Δx n ) ⎥ ⎝ ⎠ ⎝ ⎠ ⎠ ⎝ ⎦ ⎣ ⎜
⎟
⎜
⎟
2
2
⎜
⎟
2
2
(1) 2 1/2
2
Δy ⎛ ∂f ⎞ ⎛ Δx1 ⎞ + ⎛ ∂f ⎞ ⎛ Δx2 ⎞ + ⋯⎛ ∂f ⎞ ⎛ Δx n ⎞ ⎤ =⎡ ⎥ ⎢ ∂x1 y ⎠ ⎝ y ⎠ ⎝ ∂x2 ⎠ ⎝ y ⎠ ⎝ ∂x n ⎠ ⎝ y ⎠ ⎦ ⎣⎝ ⎜
⎟
⎜
⎟
⎜
⎟
⎜
⎟
⎜
⎟
⎜
⎟
(2)
where, x i was the independent variable, Δx i was the absolute uncertainty associated with the variable x i , y was the dependent variable, Δy was its absolute uncertainty, and f was a function of the independent variable x i . For this study, the main parameters used to evaluate the performance of the internally cooled desiccant wheels were the outlet temperature and humidity ratio. They were measured by Class B resistance temperature detectors and dew-point temperature monitors, respectively. The accuracy details for these two kinds of sensors were 1.2 °C and 0.2 °C [16], respectively. Eqs. (1) and (2) were then used to calculate the relative uncertainties, which were approximately ± 12% and ± 20% for temperature and humidity, respectively.
Fig. 6. Humidity difference between inlet and outlet process air for different air face velocities and rotation speeds.
and face velocity for supply/regeneration air, and rotation speed of the wheel) are set out in Table 1. Other test conditions had the process air and regeneration air inlets’ absolute humidity set to the same measurements. The relative humidity ratios for the process air tested were 20%, 40%, 60% and 90%. The inlet temperature for cooling water was set to 15 °C to –30 °C and flow rate was set to 0–3.5 L/min (also termed the ‘dry performance’ for the new wheel whenever the water flow rate was 0 L/min). This relatively low water flow range was caused by how the wheel was constructed. Given that the four wheel sectors were 3D printed separately, and did not fit each other precisely, it was necessary to use a thin waterproof sealant (silastic) between each sector. It was then discovered that, if the water flow was too high, the high water flow could damage the waterproofing used to seal the porous structure.
2.4. Description for experimental test conditions The wheel was tested over a range of process and regeneration air inlet conditions designed to cover the typical range of operating points, as suggested by the manufacturer [20] of the conventional desiccant wheel, which covered an inlet process air temperature range of 25 °C–35 °C. Some test condition combinations (including temperature
Fig. 5. Schematic chart for the cooling loop. 5
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Fig. 7. Thermal process change of the process air in psychrometric chart when a) inlet air temperature was 25 °C and inlet regeneration air temperature was 50 °C, b) 25 °C and 60 °C, and c) 25 °C and 70 °C. (Cooling water temperature was 24 °C and flow rate was 1.5 L/min for all three figures.)
3. Experiment results
are shown in Fig. 6. Note that, in Fig. 6, there is no recorded data for 30 rph of rotation speed and face velocity of 2.5 m/s. This was because the experimental system failed to maintain a steady state during this test. The maximum error of the outlet reading was about 1.6 g/kg (about ± 0.8 g/kg to the average value), which was far greater than the uncertainty of the typical values obtained from the experimental system. Other findings include that, when the rotation speed of the wheel was fixed (at 15, 20, 25 and 30 rph), the humidity difference between the inlet and outlet process air increased first when air face velocity varied from 1.5 to 2.0 m/s, then decreased when air face velocity increased to 3.0 m/s. The maximum difference happened in the range of 1.5–2.0 m/s air face velocity. When the air face velocity was constant, in most cases the humidity difference for 20 and 25 rph was larger than that for 15 and 30 rph. The largest humidity difference occurred when the air face velocity was 2.0 m/s for the selected wheel rotation speeds (15, 20, 25 and 30 rph). When the rotation speed was 20 rph, the average dehumidification performance was better than in the conditions with different rotation speeds. Thus, 2.0 m/s face velocity and 20 rph rotation speed were selected to do the remainder of the performance tests for the new wheel, for which the optimised rotation speed and face velocity when the internally cooled desiccant both had a relatively low process air outlet temperature and humidity (10–25 rph and 1.0–2.0 m/s).
To investigate the overall performance of the internal heat-exchange desiccant wheel, this section analyses the performance of the desiccant wheel under different conditions: face velocity of air streams, inlet temperature and humidity of air streams, and rotation speeds of desiccant wheel, for both dry and wet performance. This work will also implement a comparison between conventional desiccant wheels and internally cooled desiccant wheels to clarify the dehumidification improvement by changing from an adiabatic dehumidification process to an isothermal dehumidification process.
3.1. Analysis of results for various face velocity of air streams and rotation speeds of the concentric design wheel To ensure that the concentric design wheel could work in the optimal air face velocity and rotation speed range, the optimal air face velocity and rotation speed of the wheel was determined before the series of tests began. Here, the optimal air face velocity and rotation speed were those in which the desiccant wheel had the maximum humidity difference between the process air inlet and outlet. The test conditions for these investigations were as follows: process and regeneration air temperatures were between 30 °C and 50 °C, respectively; air face velocity varied from 1.5 to 3.0 m/s (process and regeneration air had the same face velocity); wheel’s rotation speed varied from 15 to 30 rph (resolution per hour); temperature for cooling water was 24 °C; and flow rate was 1.5 L/min. The experiment results 6
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Fig. 8. Thermal process change of the process air in psychrometric chart when a) inlet air temperature was 35 °C and inlet regeneration air temperature is 50 °C, b) 35 °C and 60 °C, and c) 35 °C and 70 °C. (Cooling water temperature was 24 °C and flow rate was 1.5 L/min for all three figures.)
(for wet performance) than the outlet temperature for a conventional desiccant wheel. The small temperature difference between the outlet temperatures for the wet performance of the new wheel and the performance for a conventional desiccant wheel was caused by the tiny temperature difference of the cooling water (24 °C) and process air inlet’s temperature (25 °C). For the process air outlet’s humidity, the results for dry performance was, on average, 1.1 g/kg higher than that of conventional performance, while the results for wet performance was, on average, 0.6 g/kg higher than that of conventional performance. The process air outlet’s humidity in wet performance fluctuated around the results of conventional performance. This meant that the small temperature difference between the cooling water and process air inlet did not influence the dehumidification performance of the wheel too much when the process air inlet’s temperature was 25 °C. However, for the process air’s outlet temperature, when the process air inlet’s temperature was 35 °C, the process air outlet’s temperature in dry performance was also slightly higher (around 0.6 in average) than that of conventional performance while the process air temperature in wet performance was, on average, 7 °C lower than that of conventional performance. This meant that, when the temperature difference between the cooling water (24 °C) and the process air inlet (35 °C) became larger, the thermal performance (regarding the process air outlet’s temperature) became better. The process air outlet’s humidity, for dry performance, was 0.4 g/kg higher than that for conventional performance while the outlet humidity for wet performance was, on average, 0.4 g/kg lower than that
3.2. Analysis of results for various inlet temperature and humidity of air streams Fig. 7(a), (b) and (c) show the changing thermal conditions of the process air in a psychrometric chart, when the process and regeneration air inlet temperatures were 25 °C/50 °C, 25 °C/60 °C and 25 °C/70 °C, respectively. Fig. 8(a), (b) and (c) show the changing thermal conditions of the process air in a psychrometric chart, when the process air/ regeneration air inlet temperatures were 35 °C/50 °C, 35 °C/60 °C and 35 °C/70 °C for conventional desiccant wheels (green rotor wheels with a quasi-adiabatic adiabatic dehumidification process using results data provided by the manufacturer [20]) in both the dry performance for the new wheel (with the test results for the new wheel when it worked while cooling water did not flow inside it) and the wet performance for the concentric structure-based wheel (with the test results when it worked while cooling water flowed inside it). The desiccant materials used in these wheels were exactly same, including their isothermal properties and channel shapes. The results in Figs. 7 and 8 show that, when the inlet temperature of process air was 25 °C, the process air outlet’s temperature in conventional performance was, on average, 2.4 °C lower than that of the dry performance, while the process air outlet’s temperature in the wet performance was, on average, 3 °C lower than that of conventional performance. This implies two things: the first was that the extra material used to construct the new wheel did not negatively impact the performance of the wheel too much due to its thermal capacity. The second was that the new wheel could have a lower outlet temperature 7
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Fig. 9. Dehumidification performance for dry and wet performance of the concentric design wheel when inlet process air temperature was a) 25 °C and b) 35 °C.
Angrisani et al. [19] in 2012. ii. The dehumidification performance for the wet process was always better than for the dry process (even in the same inlet conditions), and the average increase was about 0.11 kg/h.
of conventional performance. However, the theoretical humidity difference between the process air inlet and outlet from the simulated results after optimisation in [16] was more than 1.4 g/kg. Thus, the improvement of dehumidification performance (in wet performance) for the new wheel was far from ideal. The potential reasons for this will be explained in the discussion section.
3.4. Performance improvement analysis for the concentric design wheel based on the enthalpy effectiveness
3.3. Analysis of results for dehumidification performance of concentric design wheel in dry and wet performance
Both the process air outlet’s temperature and humidity need to be considered when evaluating the performance improvement of the desiccant wheel, changing from an adiabatic dehumidification process to an isothermal dehumidification process. Thus, a new evaluation standard proposed by Mandageri [21] was adopted to evaluate the overall performance improvement of the concentric design wheel compared to a conventional desiccant wheel. This new standard was termed the ‘enthalpy effectiveness’ of the desiccant wheel and showed the deviation of the outlet process air enthalpy from the inlet air enthalpy, which accounted for both outlet temperature and humidity (defined below):
The dehumidification performance of the desiccant wheel was defined using the following equation:
D = ṁ a (dp, in − dp, out )
(3)
Fig. 9(a) and (b) show the dehumidification performance of the new wheel (both dry and wet performance) when the cooling water temperature was 24 °C, flow rate was 1.5 L/min, wheel’s rotation speed was 20 rph, air face velocity for both process air and regeneration air was 2.0 m/s, and the inlet absolute humidity ratio for both process and regeneration air was the same. Two main findings can be concluded from Fig. 9(a) and (b): i. When the temperature of both the process and regeneration air was kept constant, the concentric design-based wheel’s dehumidification performance increased as the inlet air relative humidity ratio increased. This occurred under both dry conditions (no water cooling) and wet conditions (with water cooling). When the inlet temperature of the process air was kept constant, the concentric design-based wheel’s dehumidification performance improved as the regeneration air temperature increased. This finding is in agreement with that of a conventional desiccant wheel, as reported by
η=1−
hp, out − hp, out , ideal hp, out , ideal
h = 1.01T + (2500 + 1.84T ) d
(4) (5)
Eq. (4) shows that, when the process air outlet’s enthalpy was the ideal value (this was the value when the process air outlet worked in an isothermal process, which can be calculated using the model introduced in [16]), the enthalpy effectiveness is 1. The closer the process air outlet’s enthalpy is to the ideal enthalpy, the larger enthalpy effectiveness it will have. However, with the conventional desiccant wheel, the released heat due to the adsorption of water vapour by the 8
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Fig. 10. Enthalpy effectiveness for the concentric design wheel when temperature for process/regeneration air was: a) 25/50 °C, b) 25/60 °C and c) 25/70 °C (Cooling water temperature was 24 °C and flow rate was 1.5 L/min for all three figures).
process air inlet’s humidity ratio increased.
desiccant, and the heat carried over from the regeneration section and stored in the matrix, would lead to a quasi-adiabatic dehumidification process, which provides an effectiveness of 0.35–0.84 for the superadsorbent desiccant used in this study (the effectiveness is calculated out based on the data provided by [16,20]). Figs. 10 and 11 show the enthalpy effectiveness of the concentric design wheel under varying weather conditions, for both dry and wet performance, along with the enthalpy effectiveness of the conventional desiccant wheel (as mentioned above). As shown in Fig. 10, the conventional desiccant wheel’s enthalpy effectiveness was higher than that of the new wheel when it worked in dry performance mode. This implies that the structural material used to form the cooling water and desiccant channel could create extra thermal mass due to its thermal capacity. This extra thermal mass had a negative influence on the performance of the wheel. The enthalpy effectiveness for the new wheel when it worked in wet performance mode was always slightly higher than that of the conventional wheel. This implies that the introduction of cooling water into the wheel’s process section could strengthen the wheel’s enthalpy effectiveness. However, this improvement was not obvious (an average increase of 4%), due to the small temperature difference between the cooling water and the process air. The finding concluded from Fig. 11 was that, when the process air inlet’s temperature was 35 °C, the wheel’s enthalpy effectiveness in wet performance mode was much higher than that of the conventional wheel (an average increase of 11%), but it was still lower than that of the conventional wheel when it worked in dry performance mode. When the process air inlet’s temperature was at 25 °C and 35 °C, the enthalpy effectiveness decreased as the regeneration temperature and
3.5. Performance comparisons between conventional desiccant wheels and internally cooled desiccant wheels To better understand the performance of the conventional desiccant wheel (using different materials, including zeolite, silica gel and polymer desiccant) and internally cooled desiccant wheels (with tubeshell and concentric structures) regenerated by a low regeneration temperature (50 °C), the performance result comparisons are given in Fig. 12. The setting conditions were as follows: rotation speed of all desiccant wheels was 20 rph; inlet temperature for the cooling water was 24 °C, and flow rate was 4.5 L/min for the internally cooled desiccant wheels; inlet temperature was 30 °C for the supply air and 50 °C for the regeneration air, for all desiccant wheels; and relative humidity ratio varied from 30% to 90% for the supply air flow (the regeneration air had the same absolute humidity as the supply air), for all desiccant wheels. Because conventional and internally cooled desiccant wheels both have different optimal face velocities, a face velocity of 2.5 m/s [22] was set for both supply and regeneration air flow for the conventional desiccant wheels, while a face velocity of 2.0 m/s [15,16,23] was set for both supply and regeneration air flow of the internally cooled desiccant wheels. As can be seen in Fig. 12, when using low temperature of regenerative heat, the performance of the conventional desiccant wheel using a zeolite desiccant was not as good as the other wheels, because it had the highest outlet temperature and humidity. The outlet temperature of the conventional desiccant wheel using silica gel desiccant was always higher than that using a polymer desiccant. While the outlet 9
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Fig. 11. Enthalpy effectiveness for the concentric wheel when temperature for process/regeneration air was: a) 35/50 °C, b) 35/60 °C and c) 35/70 °C (Cooling water temperature was 24 °C and flow rate was 1.5 L/min for all three figures).
0.4 g/kg) was not as good as the simulated dehumidification improvement (on average, 0.79 g/kg). There were three reasons that can explain this unexpected performance. First, the width of the channels used for cooling water was too narrow (the channels were less than 2 mm, which was narrower than the channels in the tube-shell design internally cooled desiccant wheel), due to an inaccuracy in the 3D printing technology. When the desiccant channels in the wheel’s process section rotated to the regeneration section, the cooling water did not have enough time to fall into the process section, which could then be brought to the wheel’s regeneration section and exchange heat with the regeneration air. The second reason was that, due to the fragile waterproof sealant, the flow rate of the cooling water was too low (the maximum water flow rate was only 3.5 L/min); this significantly influenced the heatexchange process between the process air and cooling water. The final reason was that the desiccant wheel had a porous structure due to the poor 3D printing technology, which led to water leakage. Although surface sealing work had been performed before the formal test, there was still a small amount of water leakage (which could increase the air humidity ratio about 1.2 g/kg) from water channels into the desiccant channels, which could weaken the new wheel’s dehumidification performance. To improve the performance of the internally cooled desiccant wheel, a more accurate manufacturing process is required to produce the concentric structure. In addition, the use of nonporous materials and a simpler wheel structure would also be advantageous. These approaches are recommended as future work, which would need to be done to move the technology towards a commercial product.
humidity of the conventional desiccant wheel using silica gel desiccant was higher than that using polymer desiccant, this was only when the humidity was high (i.e., relative humidity ratio greater than 60% [24]). The performance of the internally cooled desiccant wheel with a tubeshell structure was only a small improvement to the conventional desiccant wheel using polymer desiccant, due to the high heat resistance caused by the overlap of multiple desiccant layers. However, the performance of the internally cooled desiccant wheel with a concentric structure improved significantly compared to the conventional desiccant wheel using a polymer desiccant, because the outlet temperature and humidity was more than 3 °C lower and 1 g/kg higher than those of the conventional desiccant wheel using a polymer desiccant. 4. Discussion This study investigated the cooling and dehumidification performance of a new internally cooled desiccant wheel with a counter-flow configuration between the cooling water and process air. In the new wheel, the high heat resistance caused by the overlap of too many desiccant layers was avoided. This was because, instead of wrapping many desiccant layers in a column inside a heat-exchange tube, the desiccant layers were attached inside concentric rings around the water channels. The maximum number of desiccant layers between any two water channels was only four. The ratio of the face area used for processing air was about 67% of the total face area. Another advantage of this wheel compared to existing non-adiabatic wheels was that the new wheel had a counter-flow arrangement between the process and the cooling source, which had a better heat-exchange performance than the crossflow arrangements in existing non-adiabatic wheels. However, the overall dehumidification improvement (on average, 10
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Fig. 12. Outlet temperature (a) and humidity (b) comparisons between conventional desiccant wheels (using zeolite, silica gel and polymer desiccant [22]) and internally cooled desiccant wheels (based on tube-shell and concentric designs).
5. Conclusion
The new wheel design uses the lowest temperature (50 °C–70 °C) of thermal energy possible, including solar thermal and waste heat. Thus, the next stage of work will be to analyse the feasibility for the application of the desiccant wheel-based cooling system using the internally cooled wheel in various climate areas.
In this study, an internally cooled desiccant wheel-based on a concentric structure was designed, constructed, tested and analysed. Compared to the tube-shell design [16], this design did not have multiple desiccant layers overlapping each other, so the high heat resistance could be avoided. The water flow in this design had a counterflow arrangement with the process air; this had a better heat-exchange performance than that present in the tube-shell design. The new wheel’s performance was tested for various weather conditions; the results were plotted in a psychrometric chart and compared to a conventional desiccant wheel’s performance to evaluate its dehumidification process. During the evaluation, enthalpy effectiveness was applied to analyse the non-adiabatic performance of both wheels. The main findings can be summarised as follows:
Declaration of Competing Interest The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper. Acknowledgements This project is part of the desiccant cooling program conducted by the University of New South Wales and CSIRO and funded by the CRC for Low Carbon Living Ltd, which is supported by the Cooperative Research Centres program, an Australian Government initiative.
1. When the temperature difference between the process air (25 °C) and cooling water (24 °C) was too small, the thermal performance of the concentric design wheel did not improve significantly. When this difference became larger (24 °C for cooling water and 35 °C for process air), the thermal process of the new wheel improved significantly, shifting from an adiabatic process to a semi-isothermal process. 2. The concentric design wheel’s enthalpy effectiveness decreased with an increase in the process air inlet’s relative humidity and the regeneration air temperature. 3. The concentric design wheel’s enthalpy effectiveness was higher than that of the conventional desiccant wheel (on average, 11% and 4% higher for process air inlet temperatures of 35 °C and 25 °C, respectively) when it worked in wet performance mode. It was lower (on average, 3% and 13% lower for process air inlet temperatures of 35 °C and 25 °C, respectively) than that of the conventional desiccant wheel when it worked in dry performance mode.
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