Application Potential of a Dew-Point Cooling Tower in Selected Energy Intensive Applications in Temperate Climate

Abstract

In the article, the application potential of the dew-point cooling tower (DPCT) in selected energy-intensive applications in temperate climates was analyzed and discussed. The applications selected for analysis are power generation with natural gas turbines and chilled water air conditioning systems. The study is based on a mathematical model derived from a modified ε-NTU model. The model was validated against experimental results and showed satisfactory agreement with the experimental data. DPCT was compared with a typical cooling tower limited by the wet-bulb temperature (wet-bulb cooling tower, WBCT). The simulation results showed that DPCT is able to provide significant energy savings in energy-intensive applications; therefore, its application potential in temperate climates can be considered justified. In the case of gas turbines, DPCT was able to generate 2 to 10 percentage points more capacity than operating on outdoor air and 1.8 to 5 percentage points more than operating with WBCT. In the case of air conditioning systems, the system equipped with DPCT achieved EERs (energy efficiency ratios) higher by 1 to 7.2 compared to dry cooling and by 0.3 to 5.1 compared to systems equipped with WBCT. The annual energy savings obtained by the system with DPCT were 14.7 MWh compared to WBCT and 30 MWh compared to dry cooling.

1. Introduction

Cooling towers play a crucial role in various industrial processes and HVAC (heating, ventilation, and air conditioning) systems by dissipating heat from water-cooling systems. They are essential in power plants, refineries, chemical plants, and manufacturing facilities where large amounts of heat are generated during their operation [1]. Cooling towers help regulate the temperature of the water used in industrial processes, preventing equipment from overheating and maintaining optimal operating conditions. By efficiently removing excess heat from the water, cooling towers contribute to energy efficiency and reduce the risk of equipment failure. Cooling towers also help mitigate environmental impacts by reducing thermal pollution in nearby water bodies and minimizing the reliance on freshwater sources for cooling purposes. Overall, cooling towers are indispensable components of industrial infrastructure, ensuring the safe and efficient operation of various industrial processes while promoting sustainability and environmental stewardship [2].

Many authors have been working on analyzing and improving cooling towers in the past. Wang et al. [3] explored ways to improve the performance of mechanical draft cooling towers by using various deflector plates. They experimented with four different deflector plates to achieve a more uniform flow distribution within the towers. The study also analyzed the impact of varying water flow rates and fan speeds on thermal performance. It was concluded that the fourth case study provided optimal performance, as evidenced by a reduction in water temperature by 0.22 °C and a 1.53% improvement in efficiency. Additionally, this case saw a 6.35% increase in the Merkel number, indicating superior thermal performance compared to the original configuration, especially when subjected to variations in water flow and fan speed. Ma et al. [4] presented a thermo-economic analysis of the influence of adjacent dry cooling towers on the energy generation capabilities of thermal power units. Their study involved a thermo-hydraulic analysis conducted using advanced computational software to assess the cooling towers and the power unit’s effectiveness. The findings revealed that the interaction between cooling towers is primarily influenced by heat rejection from radiators located in neighboring halves. Furthermore, an increase in the unit load was shown to enhance the performance of both cooling towers when subjected to crosswind conditions, indicating the significant impact of operational load on cooling efficiency. Arefimanesh and Heyhat [5] presented a thermodynamic model to analyze the water consumption and cooling performance of wet cooling towers under various weather conditions, validated with numerical and experimental data. The model accounts for buoyancy force, droplet evaporation, and fouling effects, using an optimization algorithm for accurate results. Applied to small cooling towers in different Iranian cities during peak summer, findings show that increased fouling decreases the fill performance index, raising outlet water temperatures while reducing water consumption. The study highlights that ambient conditions and fouling significantly impact cooling tower performance, as effectively predicted by the new model. Wang et al. [6] focused on the aspect of water conservation within mechanical draft cooling towers. They developed a sophisticated multiphase flow model to simulate the condensation process from wet air and to evaluate the water-saving system’s efficiency. The results indicated that the optimal length of the channel for maximum water-saving efficiency corresponds to a relative length of 0.6. Additionally, it was observed that reducing the inlet airflow temperature by 12 K led to a decrease in the outlet humidity from the cooling tower by 1 g/kg, showcasing a clear relationship between inlet conditions and humidity levels. Zhou et al. [6] performed a study to enhance the heat transfer performance of closed-circuit cooling towers by examining elliptical-tube heat exchangers and introducing an optimized deflected elliptical tube design. Through a combination of experiments and simulations, the researchers compared the heat transfer performance of these tubes under various conditions, including different wind speeds and spray water flow rates. The findings reveal that, at wind speeds of 2–3.2 m/s, the optimized deflected elliptical tube bundle increases the heat transfer coefficient by 16.27% due to closer tube spacing. Moreover, these optimized tubes not only outperform the standard elliptical and deflected elliptical tubes but also reduce the overall unit structure by 24.03%, leading to lower material usage in the cooling tower. Al-Qazzaza et al. [7] explored the advantages of using an underground heat exchanger instead of a wet cooling tower, particularly in terms of energy efficiency and environmental benefits. The research focuses on Mashhad, a region with high cooling demands and limited water resources, using the Faculty of Engineering at Ferdowsi University of Mashhad as a case study. Both cooling systems were modeled and simulated to assess energy and water usage. Findings indicate that the underground heat exchanger cuts energy consumption by 14% and completely eliminates water usage, which is significant given the current water-cooled chillers’ high water consumption. Additionally, the economic analysis shows that the underground heat exchanger has a payback period of 1.73 years, demonstrating its feasibility as an alternative cooling method. It can be seen that many authors have researched the subject of improving the performance of cooling towers.

One of the best methods to improve the application potential of cooling towers is to use the dew-point cooling process [8,9]. Dew-point cooling allows significantly extended water cooling abilities by moving their physical limit from the wet-bulb temperature to the dew-point temperature [8,9]. The structure of the dew-point cooling tower (DPCT) compared to the traditional wet-bulb cooling tower (WBCT) is presented in Figure 1. WBCT has all channels wetted with water; due to this fact, its effectiveness is limited by the wet-bulb temperature of the airstream at the inlet to the wet-channel (1_i in Figure 1a,c). DPCT uses a different plate heat exchanger structure, where half of the plates are covered with cooled water and half of the plates remain dry. The air flows through dry channels first, where it is pre-cooled without change in its humidity ratio (1_i = 1_o in Figure 1b,c). After that, it is delivered to the wet channel, where it cools the process water. Due to the pre-cooling effect (cooling the airstream without changing its humidity ratio), the wet-bulb temperature of the airflow decreases (reaching values close to the dew point temperature [8,9]). Due to this fact, the physical limit of the cooling process is the wet-bulb temperature of the airflow at the end of the dry channel, which enters the wet channel (1_o = 2_i). Since the wet-bulb temperature of the airstream is much lower, DPCT is able to achieve much lower temperatures than WBCT.

The authors have performed a study where they analyzed the potential of applying DPCT and WBCT in different climate conditions [9]. DPCT achieves the highest performance factors in cold and dry climates; therefore, from the economic standpoint of the tower itself, it has the most advantages over WBCT in such climate conditions. However, it can make the most difference in hot and moist climates, where it can visibly outperform the traditional towers, which may be more beneficial for the final user. It was established that DPCT is highly suitable for warm arid and desert climates, especially hot semi-arid climates, hot and cold desert climates, and hot and warm continental climates with dry summers (e.g., Poland). These conditions can be considered optimal for the operation of the DPCT.

Lower water temperatures allow for more efficient heat rejection and, therefore, lower energy consumption of the supplied equipment (e.g., power plants, air conditioning systems, and refrigeration equipment). So far, no studies have focused on analyzing the application potential of dew-point cooling towers in energy-intensive applications in order to establish what level of energy savings can be expected when traditional WBCTs are replaced by DPCTs. Due to this fact, the proposed study is focused on analyzing the application potential of the dew-point cooling tower in selected energy-intensive applications in temperate climates (using the example of Poland). The applications chosen for the proposed analysis are gas power plants and air conditioning systems. The applications were selected due to their high energy consumption and the fact that they are critical for the economy: air conditioning and refrigeration systems are necessary to provide acceptable indoor conditions during the summer season, and power plants are necessary to produce power for every sector of the country. The ability to reduce their energy consumption is critical and highly beneficial to society. The analysis will be performed using experimentally validated mathematical models.

2. Methods

The study is based on simplified assumptions in order to compare the theoretical potential of the dew-point cooling tower concept. Due to the different nature of the structures of DPCT and WBCT, the pressure drop analysis was omitted, along with the energy consumption of the supplementary equipment (pumps, fans, additional heat exchangers, etc.). This is caused by the fact that in real-life situations they would use a different construction; therefore, it would be very hard to present an objective comparison between two completely different structures. The goal of the paper is to analyze their theoretical potential; therefore, only the thermodynamic performance will be used as a factor in the comparison.

2.1. Numerical Model

For the purpose of analysis, a numerical model based on a modified ε–NTU method was implemented. The proposed model assumes that the airstream is considered a gaseous fluid with constant mass transfer potential, temperature, and velocity which, in sections normal to the exchanger plate, are equal to their bulk average values [10,11]. The assumptions for the model include the state, the treatment of air as an ideal gas, and the fact that the humidity ratio gradient is the main mechanism of mass transfer.

The processes of heat and mass transfer are described using ordinary differential equations, where the stream in the dry channel is marked as 1, the stream in the wet channel is marked as 2, and the water stream is marked as W (Figure 2). The details of the numerical model used in this study were presented by the authors in [8,9], therefore they will be omitted in the current paper; only the basic equations will be presented.

The equation for the sensible heat balance for the airstream in the dry channel is given as

(1)${\displaystyle \frac{d{t}_1}{d\overline{X}}}={\mathrm{NTU}}_1^{\prime }\left(t_{p1}-t_1\right)$

The equation for the mass balance for the airstream in the wet channel is given as

(2)${\displaystyle \frac{d{x}_2}{d\overline{X}}}=-{\mathrm{NTU}}_2^{\prime }\left({\displaystyle \frac{1}{{\mathrm{Le}}_2}}\right)\left({x^{\prime }}_w-x_2\right)$

The equation for the sensible heat balance for the airstream in the wet channel is given as

(3)${\displaystyle \frac{d{t}_2}{d\overline{X}}}=-{\mathrm{NTU}}_2^{\prime }\left({t^{\prime }}_w-t_2\right)$

The equation for the mass balance for the water stream in the wet channel is given as

(4)${\displaystyle \frac{d{G}_w}{d\overline{X}}}=G_2{\mathrm{NTU}}_2^{\prime }\left({\displaystyle \frac{1}{{\mathrm{Le}}_2}}\right)\left(x_w^{\prime }-x_2\right)$

The equation for the energy balance for the water stream in the wet channel is given as

(5)${\displaystyle \frac{d{t}_w}{d\overline{X}}}=W_1/W_w{\mathrm{NTU}}_1^{\prime }\left(t_{p1}-t_1\right)+W_2/W_w{\mathrm{NTU}}_2^{\prime }\left[\left(t_w-t_2\right)+\left({\displaystyle \frac{r}{c_{p2}}}\right)\left(x_{\prime }^w-x_2\right)\right]$

where:

$-W_1/W_w=\left(G_1{c}_{p1}\right)/\left(G_w{c}_w\right)$

$-W_2/W_w=\left(G_2{c}_{p2}\right)/\left(G_w{c}_w\right);$

The convective heat transfer coefficient for the water stream in the wet channel is obtained from the Nusselt number [12]:

(6)${\mathrm{Nu}}_w={\displaystyle \frac{{\alpha }_w{\delta }_w}{{\lambda }_w}}=1.88$

The thickness of the water film is determined using the equation below [13]:

(7)${\delta }_w={\left({\displaystyle \frac{3{\nu }_w{\mathsf{\Gamma }}_w}{{\rho }_{w}g}}\right)}^{1/3}$

The density of water sprayed on the wet plates is obtained using the fusing correlation presented below [13]:

(8)${\mathsf{\Gamma }}_w={\displaystyle \frac{G_w}{\left(n+1\right)L_X}}$

2.2. Effectiveness Dependencies

For the purpose of analyzing energy-intensive applications (i.e., gas power plants and air conditioning systems), a simplified method is applied: the performance of the systems will be simulated using effectiveness curves. These methods were proven to be effective for the analysis of these types of systems in multiple studies [14,15,16].

A water-cooled chilled water unit (chiller [16]) was assumed as the example air conditioning system for the purpose of this comparison. For the purposes of analysis of the air conditioning system, an EER curve vs. supplied cooling water temperature and level of nominal capacity will be used (Figure 3). EER is defined using the following equation:

$\mathrm{EER}={\displaystyle \frac{Q_C}{N}},-$

where:

Q_C—cooling capacity obtained from the chilled water unit, W

N—power necessary to produce the required cooling capacity (for compressors and other supplementary equipment), W

Figure 3

Coefficient of energy efficiency of the water-cooled refrigeration unit as a function of performance and cooling water temperature.

[Figure omitted. See PDF]

Dew-point and wet-bulb cooling towers can be used to supply cold water to remove the heat from the condenser of the chilled water unit. Depending on the temperature of the water obtained from the tower and the percent of nominal capacity that is used by the system, the chilled water unit will obtain a different EER level, which corresponds directly to the energy consumption (the unit can supply the same capacity with lower energy consumption with a higher EER; it can also supply a higher cooling capacity with the same energy consumption when it is needed).

For the purposes of analyzing the gas power plant, a gas turbine capacity curve is going to be used (Figure 4). The curve shows the dependence of the capacity that can be obtained from the gas turbine for different air temperatures at the inlet to the compressor. The capacity is referenced to the nominal capacity for an inlet air temperature equal to 15 °C—when the supplied air is characterized by temperatures lower than 15 °C, the cooling capacity proportionally increases. When supplied air is characterized by temperatures higher than 15 °C, the capacity decreases.

Cooling towers can be used to supply the cold water to initially pre-cool the airstream before entering the compressor, which can generate energy savings. When air with a lower temperature is supplied, the gas turbine can generate the same capacity while consuming less fuel or produce more power using the same amount of fuel. Depending on the temperatures supplied by the cooling towers, the performance of the turbine will change accordingly.

2.3. Weather Data

The cooling season in Polish climate conditions lasts for app. 140 days (part of May, June, July, August, and part of September). The weather data (i.e., air temperature and relative humidity) is based on official meteorological data [17]. Temperature and humidity levels are presented in Figure 5. It can be seen that the temperatures may vary from app. 10 °C to app. 31 °C (Figure 5a), and the relative humidity may change from 29 to 95%.

3. Validation of the Mathematical Model

The model was validated against data obtained from the experiment on a laboratory-scale prototype of DPCT. The schematic of the testing station used for verification of the mathematical model is presented in Figure 6. The setup utilizes a counter-flow recuperator with half of the channels wetted with water (distributed through spray nozzles), allowing for the transfer of the pre-cooled air from dry channels to wet channels in a dew-point cycle. Water is supplied from a basin equipped with an electric heater, simulating the operational temperature of the water, which is cooled by the DPCT. A fan is used to force the air to flow through the cooling tower. Measurements were taken using Sensirion SHT85 air temperature and humidity sensors (Sensirion AG, Stäfa, Switzerland) (accuracy of 0.1 °C and 1.5% RH, respectively) and Sensirion SHT85 sensors in a water proof cover for the water temperature (accuracy of 0.1 °C). Water was supplied by a pump located in the basin. Air temperature and humidity were measured at four points in the system: at the inlet to the exchanger, directly behind the dry channel, directly before the wet channel, and at the outlet of the exchanger. Water temperature measurements were taken directly in the basins: the first basin contained warm water intended for cooling, and the second basin held already cooled fluid. To maintain continuous water flow, the cold, already-cooled water was returned to the tank with hot water. Warm water was obtained by heating it with a regular electric heater (capacity: 1.5 kW). In addition, an electric heater was used to increase the temperature of the airflow to simulate summer conditions. The inlet air temperature was stabilized at 25 °C and 10 g/kg humidity using a laboratory air conditioning system, and the water temperature varied from 50 to 60 °C with a 1 °C interval. The steady-state conditions were achieved after approximately 15 min of 45 min of each testing interval. The airflow rate was 200 m³/h, and the water flow rate was app. 90 kg/h.

The validation was performed by comparing the outlet water temperature obtained during the experimental tests and from the numerical model. The model was set on identical operational conditions as in the experimental set-up, including the exchanger geometry and inlet air and water parameters. The validation of the model (the correlation between simulation and experiment results) is presented in Figure 7. It can be observed that the highest difference in temperatures of the cooled water between the numerical and experimental results is 1.0 °C, with an average temperature difference of app. 0.5 °C (the average discrepancy, defined as the temperature difference between the inlet and outlet temperature obtained from the model divided by the temperature difference between the inlet and outlet temperature obtained from the experiment, was 5.35%). The correlation between the model and the experiment is 0.9934. Furthermore, the model and the experiment exhibit the same trend in the obtained water temperatures at the exchanger outlet, depending on the initial conditions. Therefore, it can be concluded that the agreement between experimental results and the mathematical model is satisfactory, and the model can be utilized for simulating the operation of the DPCT.

4. Results and Discussion

In this section, an analysis of the application potential of the DPCT in temperate climate conditions using the example of Poland is presented. The analysis is conducted for two examples: a gas power plant (Section 4.2) and an air conditioning system (Section 4.3).

4.1. Performance Comparison between DPCT and WBCT

First, analysis is focused on the performance comparison between wet-bulb and dew-point cooling towers in Polish climate conditions in terms of the obtainable water temperature. The results of the comparison are presented in Figure 8. It can be seen that the outlet water temperatures obtainable by DPCT vary from 4.5 °C to 21 °C, while the water temperatures obtained by the wet-bulb cooling tower vary from 7 to 26 °C. The average difference between obtainable water temperatures is 3.9 °C during the entire cooling season; the maximal obtainable difference in outlet water temperature is 5 °C; and the minimal obtainable difference is 1.0 °C. The difference in obtainable temperatures is caused by the fact that DPCT allows for the pre-cooling of the airstream without changing its humidity ratio. This allows it to decrease its wet-bulb temperature. Due to the fact that the air wet-bulb temperature at the inlet to the wet channel is the theoretical limit of the cooling tower performance, the pre-cooling process in the dry channel allows to lower the limit and, in consequence, achieve higher effectiveness. For example, for inlet air parameters equal to 30 °C and 40% RH, the wet-bulb temperature is equal to 20.1 °C, and the wet-bulb temperature of the pre-cooled primary air at the outlet of the dry channel (1_o = 2_i) is equal to 16.2 °C. This allows the water to cool to 18.3 °C, which is almost 2 °C lower than the wet-bulb temperature of the inlet air. WBCT does not have the ability to pre-cool the outdoor air; therefore, it obtains higher temperatures than DPCT. In the next two sections, the impact of the difference in obtainable water temperature on the performance of gas power plants and air conditioning systems will be investigated.

4.2. Gas Power Plants

In the case of a gas power plant, water from the cooling tower can be used for the preliminary cooling of air. The unit cools the air, thereby increasing its density, allowing for a reduction in the energy consumption for the compression process or achieving higher power with the same fuel consumption. The performance of a gas turbine supplied with air pre-cooled using water from DPCT and WBCT is presented in Figure 9. The obtainable capacity vs. nominal capacity with an inlet air temperature equal to 15 °C is presented in Figure 9a,b; the difference in obtainable capacity using the example of 600 MW gas turbine is presented in Figure 9c,d; and the potential natural gas savings are presented in Figure 9e. Maintaining a lower temperature with DPCT allows lower air temperatures at the inlet to the turbine, which corresponds to better turbine performance. The dew-point tower allows it to generate from 2 to 10 percent more capacity than on outdoor air. The wet-bulb unit is able to provide a 0.2 to 5 percent point increase in capacity. A 600 MW turbine, on average, is able to produce 44 MW more power with DPCT (8% increase on average—Figure 9c) in comparison to operating on outdoor air and over 22 MW more in comparison to the operation with WBCT (4% increase on average—Figure 9d). In Polish climatic conditions, fuel savings or excess power can be up to 10% with DPCT and up to 5% with WBCT. It is worth emphasizing that the external temperature has less impact on the DPCT performance, allowing it to provide much more stable temperatures during the cooling season, which can contribute to an additional increase in the turbine performance [18]. The potential fuel savings for a 600 MW power plant based on gas turbines vary from 1200 m³ per day to 2600 m³ per day for WBCT and from 2500 m³ per day to 5000 m³/day for WBCT. On average, the DPCT is able to save 3800 m³ of natural gas per day, whereas the WBCT is able to provide 1820 m³ of average gas savings. The cost of producing electricity from gas is approximately 160 EUR/MWh [19]. A 600 MW gas power plant would be able to save from 113 to 223 thousand EUR per day with the dew-point cooling tower and from 53 to 120 thousand EUR per day with WBCT. On average, the money savings per day would be 169 thousand EUR for DPCT and 81 thousand EUR for WBCT. Annually, the application of the dew-point cooling tower allows for 23,995,452 EUR savings, whereas the wet-bulb cooling tower allows for 11,479,039 EUR savings (the difference is 12,476,413 EUR in favor of the DPCT). It can be seen that the dew-point cooling tower is able to bring visible energy savings to a gas power plant in temperate climate conditions. The energy savings correspond to an appropriate level of money savings, which can be utilized for other purposes. It can also be seen that DPCT provides over two times more savings than the traditional solution, which shows that it has the potential to replace them in temperate climate.

4.3. Air Conditioning Systems

The performance of the air conditioning system equipped with DPCT and WBCT is presented in Figure 10. For the purpose of comparison, both systems are referenced to a chilled water system operating with an air-cooled heat rejection system (i.e., a dry cooler). The performance is analyzed for different cooling load levels: the performance under 100% of the nominal cooling load is presented in Figure 10a,b; the performance under 50% of the nominal cooling load is presented in Figure 10c,d; the performance under 30% of the nominal cooling load is presented in Figure 10e,f.

The performance of the cooling system is analyzed based on the energy efficiency ratio (EER), defined as the obtained cooling power divided by the input electrical energy (W/W), as described in Section 2. Using the obtained water temperature values and the functional relationship for the unit’s EER, the EER levels for dry heat rejection and switching to WBCT and DPCT were determined. It can be seen that during operation at 100% of the nominal cooling load, the system, while operating with dry cooling heat rejection, is able to achieve EERs varying from 3.7 to 9.8 during colder days (Figure 10a,b). When DPCT is applied, the EERs vary from 5.5 to 13.8 (Figure 10a), and when WBCT is applied, the EERs range from 4.2 to 11.1 (Figure 10b) under 100% cooling load. During reduced cooling load (50%), the EERs obtained by the system with dry cooling vary from 4.1 to 11.8 (Figure 10c,d), with DPCT achieving EERs from 6.2 to 15.9 (Figure 10c), and with WBCT achieving EERs from 5.7 to 13.0 (Figure 10d). During operation at 30% of the nominal cooling load, the EERs for dry cooling vary from 4.9 to 12.2 (Figure 10e,f); with DPCT, they range from 7.5 to 16.2, and with WBCT, they range from 5.9 to 14.0. The maximum difference in EER between the system with DPCT and dry cooling is 7.2; the minimum is 1.0; the maximum difference between the system with DPCT and WBCT is 5.1; the minimum is 0.3. The change in average EER between DPCT and dry cooling is 3.5, and between DPCT and WBCT is 2.1.

To perform a simplified analysis of the level of energy savings achieved by implementing DPCT compared to WBCT and dry cooling, a cooling system with a nominal capacity of 1 MW is assumed. The average operation scenario is based on [20,21], with average EERs, 10 h a day of air conditioning, and operation at maximum capacity occurring 15% of the time, at 50% capacity 60% of the time, and at 30% of nominal capacity 25% of the time. In such a case, a 1 MW cooling system would consume 64.5 MWh of electrical energy annually with DPCT, 79.2 MWh with WBCT, and 94.5 MWh with dry cooling. This means that the savings in electrical energy between dry cooling and DPCT would be 30 MWh, and between DPCT and WBCT would be 14.7 MWh. It can be seen that the application of DPCT can improve the performance of cooling systems and provide significant energy savings.

5. Summary and Conclusions

The proposed paper discusses the application potential of dew-point cooling towers (DPCT) in selected energy-intensive applications in temperate climates. The applications selected for analysis are power generation with natural gas turbines and chilled water air conditioning systems. The analysis was performed using a numerical model that was validated against experimental data. The results indicated that DPCT is able to provide significant energy and financial savings in energy-intensive applications; therefore, its application potential in temperate climates can be considered justified. In the case of the gas turbines, DPCT was able to generate 2 to 10 percentage points more capacity than operating on outdoor air and 1.8 to 5 percentage points more than operating with wet-bulb cooling towers (WBCT). Annually, the application of dew-point cooling towers allows for savings of approximately 12,476,413 EUR more than systems with WBCT, which corresponds to daily savings of almost 2000 cubic meters of natural gas. The maximum difference in EER (energy efficiency ratio) between systems with DPCT and dry cooling is 7.2, and the minimum is 1.0. The maximum difference between systems with DPCT and WBCT is 5.1, and the minimum difference is 0.3. The change in average EER between DPCT and dry cooling is 3.5, and between DPCT and WBCT is 2.1. This corresponds to annual energy savings of approximately 14.7 MWh compared to WBCT and 30 MWh compared to dry cooling.

It should be noted that the following study is based on simplified assumptions (i.e., no energy consumption of the cooling tower equipment, such as fans and pumps, along with thermal losses or efficiency of the intermediate heat exchangers are included). This may affect the results in real-life situations. Since both towers would probably use a different structure, an objective comparison would be hard to determine. Due to this fact, future research should include pilot projects based on DPCT prototypes and detail experimental verification of the actual level of savings in energy-intensive applications in a temperate climate.

Author Contributions

Conceptualization, D.P.; Methodology, D.P. and S.C.; Software, D.P.; Validation, D.P., M.M. and S.C.; Formal analysis, D.P. and W.W.; Investigation, D.P. and B.P.; Resources, D.P.; Data curation, D.P., M.M., P.K. and G.I.; Writing – original draft, D.P., M.M., P.K., B.P. and G.I.; Writing – review & editing, D.P., G.I., W.W. and S.C.; Visualization, D.P. and P.K.; Supervision, D.P.; Project administration, D.P. and W.W.; Funding acquisition, D.P. All authors have read and agreed to the published version of the manuscript.

Institutional Review Board Statement

Not applicable.

Informed Consent Statement

Not applicable.

Data Availability Statement

Data is contained within the article.

Conflicts of Interest

The authors declare no conflict of interest.

Nomenclature

cp	[J/(kg K)]	Specific heat capacity of moist air
cpw	[J/(kg K)]	Specific heat capacity of water vapor
cw	[J/(kg K)]	Specific heat capacity of water
dh	[m]	Hydraulic diameter
F	[m2]	Surface area
h	[m]	Height
g	[m/s2]	Gravitational acceleration
G	[kg/s]	Mass flow rate
L, l	[m]	Length
M	[kg/s]	Water vapor mass transfer rate
n	[-]	Number of channels
r	[kJ/kg]	Specific heat of water evaporation
Q	[W]	Rate of heat transfer
RH	[%]	Relative humidity
t	[°C]	Temperature
W	[W/K]	Heat capacity rate of the fluid
x	[kg/kg]	Humidity ratio
X	[m]	Coordinate along water flow direction
Y	[m]	Coordinate perpendicular to X coordinate
Special characters:
α	[W/(m2 K)]	Convective heat transfer coefficient
β	[kg/(m2 s)]	Mass transfer coefficient
δ	[m]	Thickness
λ	[W/(m K)]	Thermal conductivity
ε	[-]	Effectiveness
εWB	[-]	Wet-bulb thermal effectiveness, εWB = (twi − two)/(twi − t1WBi)
ν	[m2/s]	Kinematic viscosity
ρ	[kg/m3]	Density
σ	[-]	surface wettability factor, σ ∈ (0.0…1.0)
${\mathsf{\Gamma }}_w$	[kg/(m s)]	surface spraying density of water
Non dimensional coordinates:
Le	[-]	Lewis factor $\mathrm{L}\mathrm{e}=\alpha /\left(\beta c_p\right)$
NTU	[-]	Number of transfer units $\mathrm{N}\mathrm{T}\mathrm{U}=\alpha F/\left(G{c}_p\right)$
Nu	[-]	Nusselt number
Pr	[-]	Prandtl number
Re	[-]	Reynolds number
St	[-]	Stanton number
$\overline{X}$	[-]	$\overline{X}$ = X/lX—relative X coordinate
$\overline{Y}$	[-]	$\overline{Y}$ = Y/lY—relative Y coordinate
Subscripts:
1		Primary airflow in dry channel
2		Working airflow in wet channel
cond		Heat transfer by thermal conduction
Icond		Referenced to the first-order boundary conditions of heat transfer
IIcond		Referenced to the second-order boundary conditions of heat transfer
h		Referred to the channel height
heat		Heat transfer
i		Inlet
l		Latent heat
mass		Mass transfer
o		Outlet
p		Channel plate
s		Sensible heat
sat		Saturation state
w		Water
WB		Wet-bulb temperature
X		Respected to X coordinate
Y		Respected to Y coordinate
∙		Referenced to the elementary surface
′		Conditions at air/water interface
″		Referenced to the plate surface

Footnotes

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Figures

Figure 1. WBCT and DPCT. (a) Wet-bulb cooling tower. (b) Dew-point cooling tower. (c) Psychrometric chart.

Figure 2. Schematic of the mathematical of the DPCT.

Figure 4. Gas turbine capacity curve vs. inlet air temperature.

Figure 5. Outdoor air parameters during cooling season in Polish climate. (a) Air temperature. (b) Air relative humidity.

Figure 6. Scheme and photograph of the measuring station. (a) Scheme (nomenclature: T—water temperature sensor; T, RH—air temperature and humidity sensor). (b) Photograph.

Figure 7. Validation results: correlation between model and the experiment.

Figure 8. Obtainable outlet water temperatures in Polish climate. (a) DPCT. (b) WBCT.

Figure 9. Capacity obtainable by a gas power plant with air pre-cooling using water from different cooling towers. (a) Relative capacity—DPCT and outdoor air. (b) Relative capacity—WBCT and outdoor air. (c) Capacity obtainable by 600 MW power plant—DPCT and outdoor air. (d) Capacity obtainable by 600 MW power plant-WBCT and outdoor air. (e) Natural gas consumption: DPCT and WBCT.

Figure 9. Capacity obtainable by a gas power plant with air pre-cooling using water from different cooling towers. (a) Relative capacity—DPCT and outdoor air. (b) Relative capacity—WBCT and outdoor air. (c) Capacity obtainable by 600 MW power plant—DPCT and outdoor air. (d) Capacity obtainable by 600 MW power plant-WBCT and outdoor air. (e) Natural gas consumption: DPCT and WBCT.

Figure 10. EER of the air conditioning system with DPCT and WBCT. (a) operation under 100% of nominal capacity—DPCT and outdoor air. (b) operation under 100% of nominal capacity—WBCT and outdoor air. (c) operation under 50% of nominal capacity—DPCT and outdoor air. (d) operation under 50% of nominal capacity—WBCT and outdoor air. (e) operation under 30% of nominal capacity—DPCT and outdoor air. (f) operation under 30% of nominal capacity—WBCT and outdoor air.

Figure 10. EER of the air conditioning system with DPCT and WBCT. (a) operation under 100% of nominal capacity—DPCT and outdoor air. (b) operation under 100% of nominal capacity—WBCT and outdoor air. (c) operation under 50% of nominal capacity—DPCT and outdoor air. (d) operation under 50% of nominal capacity—WBCT and outdoor air. (e) operation under 30% of nominal capacity—DPCT and outdoor air. (f) operation under 30% of nominal capacity—WBCT and outdoor air.

Figure 10. EER of the air conditioning system with DPCT and WBCT. (a) operation under 100% of nominal capacity—DPCT and outdoor air. (b) operation under 100% of nominal capacity—WBCT and outdoor air. (c) operation under 50% of nominal capacity—DPCT and outdoor air. (d) operation under 50% of nominal capacity—WBCT and outdoor air. (e) operation under 30% of nominal capacity—DPCT and outdoor air. (f) operation under 30% of nominal capacity—WBCT and outdoor air.

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