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Abstract

This article deals with an experimental investigation of coinstantaneous heat and mass transfer phenomena between water and air in a counterflow wet cooling tower filled with a new type packing named “FCP-08.” The packing consisted of foamed ceramic corrugated board with sine waves, surface retention groove is 1.0 m high, and it has a cross-sectional test area of 0.68 m × 0.68 m. This investigation is focused mainly on the effect of the water–air mass flow ratio on the heat and mass transfer characteristics of the cooling tower, for different inlet water temperatures. The results show that the cooling water range, $R$ , and the cooling tower efficiency, $e$ , decrease with the increase in water–air mass flow ratio, $L / G$ . Meanwhile, the heat rejected by the cooling tower, $Q_{w}$ , increases with the increase in the water mass flow rate. The cooling characteristic coefficient, $K α V / L$ , slightly decreases with the increase in water–air mass flow ratio and the value is obviously higher than that of other packing investigated before. The relationship between the cooling characteristic coefficient and water–air mass flow ratio is obtained by linear fitting. The comparison between the obtained results and those found in the literature for other types of packing indicates that cooling performance of the tower with foam ceramic packing is better.

Keywords

Cooling tower heat and mass transfer characteristics foam ceramic packing water–air mass flow ratio cooling performance

Introduction

Counterflow wet cooling towers are widely used in power, chemical, and other fields as important heat exchange equipment; it is used to cool high-temperature water or other medium (here referred to cooling medium) from condensers and process equipment. The cooling medium is pressurized by cooling water pump and sent to the distribution pipe of the cooling tower and then sprayed to the water drenching filler through the nozzle. At the same time, the ambient air is drawn from the bottom of the tower and then passed through the packing layer; due to the pumping action of the tower, the air passes through and the heat exchange between the packing layer and cooling medium is fully carried out. In counterflow cooling towers, the hot water is sprayed into the air stream. The enthalpy of water decreases during the heat and mass transfer, while that of air increases.¹ The efficiency of the cooling tower is closely related to the heat and mass transfer characteristics.

The packing may be qualified as the heart of the cooling tower, and its thermal performance has an important influence on the heat transfer efficiency of the cooling tower. The material of packing has great influence on its thermal performance as it provides a very large surface area for evaporative heat and mass transfer to take place from hot water to the surrounding air and increases the contact time between the two fluids. In this respect, with the development of materials science and manufacturing process, various types of packing have been designed over the years and the researches have been pursued in this field in order to improve the transport phenomena in these equipments and therefore enhance their performances.

Since the early 40s, Simpson and Sherwood² studied the working performance of mechanical ventilation cooling tower with 1.05 m wood slats packing. They give data for several different tower designs over a range of performance conditions. These data were later correlated by Braun³ in terms of the number of transfer units (NTUs) as a function of the water–air mass flow ratio $(L / G)$ for the different cooling tower designs. The hydraulic and thermodynamic characteristics of three-phase fluidized bed cooling tower were studied theoretically and experimentally by EL-Dessouky.⁴ He used spongy rubber balls with a diameter of 12.7 mm and a density of 375 kg/m³ as a packing and analyzed the relationship between the tower characteristic, inlet water temperature, packing layer height, and the water–air mass flux ratio. Bedekar et al.⁵ studied the performance of a counterflow mechanical ventilation cooling tower equipped with thin film–type water spray packing by means of experiments. The performance of the cooling tower, water outlet temperature, and efficiency is expressed as functions of the water-to-air flow rate ratio, $L / G$ . They considered that the tower performance decreases with an increase in the $L / G$ ratio, but the correlation was not derived. Gharagheizi et al.¹ experimented with vertical corrugated packing (VCP) and horizontal corrugated packing (HCP) made from reinforced polyvinyl chloride (PVC) in a count flow induced draft mechanical cooling tower (IDCT) and considered that the water–air mass flow ratio, the type of packing, and its arrangement method have effect on the heat transfer performance of the cooling tower. Recently, Goshayshi et al.⁶ studied the performance of various PVC plastic packing by testing the mass transfer coefficient and pressure loss. They experimented in a cooling tower and introduced a new packing comparison method to obtain the best packing in geometric structure. They⁶ thought that the mass transfer coefficient of the corrugated packing is proportional to the pressure loss of unit height. Goshayshi and Missenden⁷ also studied the mass transfer and pressure loss characteristics of various corrugated packing (smooth or rough) in atmospheric cooling tower. Based on the experimental data, the expression of heat and mass transfer characteristics and pressure loss was derived. More recently, Elsarrag⁸ presented an experimental study and predictions of an induced draft ceramic tile packing cooling tower. He used a tower of 0.64 m² cross-sectional area and 2 m height with a filling portion of 0.8 m. He pointed out that the factors affecting the heat and mass transfer coefficients are the water-to-air flow rate ratio, the inlet water temperature, and the inlet air enthalpy. Lemouari et al.⁹ studied the effects of different air and cooling water rates on the global heat and mass transfer coefficients and water evaporation rate at different inlet temperatures in a forced counterflow cooling tower with vertical grid apparatus (VGA) type packing experimentally. The packing consists of four pieces of galvanized “Z”-shaped flakes, 0.42 m high, and the cross-sectional area of the space between the grids is 0.022 m². Their study showed that the global heat and mass transfer coefficient and water evaporation rate vary linearly with the flow rate of air and cooling water. In addition, Lemouari and Boumaza¹⁰ also studied the total heat transfer and the thermal efficiency in the forced counterflow cooling tower using VGA type packing experimentally, considering the influence of air flow rate, cooling water flow rate, and inlet temperature of cooling water. The results show that the total heat transfer and the thermal efficiency of the cooling tower increase with the increase in inlet temperature. The total heat transfer capacity increases with the increase in air flow rate and cooling water flow rate, but the cooling efficiency is opposite.

The main purpose of this article is to carry out an experimental investigation of the heat and mass transfer phenomena observed inside a counterflow wet cooling tower filled with an “FCP-08”-type packing. This type of packing which has been initially developed and used in catalyst carriers, porous media combustion, biofilters, and biotrickling filters (reported in the literature^11,12) has not been used in cooling water systems using direct contact between water and air. Based on the above considerations, this article analyzes the cooling water range, the efficiency of the cooling tower, the cooling characteristic coefficient, and the heat rejected by the cooling tower with the change in the inlet water temperature and the water–air mass flow ratio, which can provide reference for the engineering design and the application of foam ceramic packing cooling tower.

Experimental materials and devices

Packing materials and structure

Packing materials

This experiment adopts the free sintering foam ceramic packing, which uses the method of foam plastic precursor slurry forming, made by curing at room temperature. Its aggregate is powdered aluminosilicate minerals, with a fineness of 325 mesh. The main suspending agent is sodium bentonite, and the auxiliary suspending agent is carboxymethyl cellulose (CMC); the content of Na is 6.5%–8.5%. The surfactant is JFC (Fatty Alcohol-polyoxyethylene Ethe), the superplasticizer is polycarboxylic acid polymer compound, and the foam plastic carrier is polyurethane plastic foam with an aperture of 8 ppi (i.e. the number of holes per inch or 25.4 mm length); its main ingredient and physical properties are shown in Table 1. Its morphology is shown in Figure 1. The surface of foam ceramic packing was pretreated by applying the water-absorbent resin coating evenly to improve the hygroscopicity and water dispersion of the foam ceramic packing.

Table 1.

The main ingredients (mass fraction) and physical properties of foam ceramic packing.

Aggregate	Sodium bentonite	CMC	JFC	Water reducer	Binder	Water
100	3.5	0.5	0.5	2.0	25	20
Aperture (ppi)	Rate of open hole (%)	Skeleton diameter (mm)		Apparent density (g/m³)	Normal temperature compressive strength (MPa)
8	81–86	0.65–1.05		0.26–0.31		0.75–1.05

CMC: carboxymethyl cellulose.

Figure 1.

Foam ceramic packing’s morphology: (a) the SEM microstructure of the filler and (b) three-dimensional connected hole network structure of the filler.

Packing structure

The hydrodynamics and heat and mass transfer performance of the packing are closely related to their structures. Foam ceramic corrugated board has three-dimensional connected hole network structure composed of sine wave. The height of the wave is 35 mm. The wave distance is 60 mm, and the inclination of the wave is 50°. Combined with the shape of mature structured packing, we fabricated the “FCP-08” ceramic foam corrugated board packing with binder using the assembly method of sheet layer and layer crossing. The average diameter of the inclined through-hole is 45 mm, as shown in Figure 2. In order to improve the specific surface area of packing and the flow path of water membrane, and prolong the exchange time of the heat and mass, the inner and outer walls of the corrugated plate are pressed to have striated retention grooves. The retention groove is perpendicular to the height direction of the through-hole, and the average depth of the retention groove is one-third of the thickness of the packing.

Figure 2.

FCP-08 packing block’s structure: (a) inclined through-hole structure of the packing and (b) surface structure of foam ceramic corrugated board.

The structural characteristics of the packing are characterized by specific surface area and effective porosity. The specific surface area refers to the surface area of the packing per unit volume, and the effective porosity refers to the total volume of voids that are interconnected in the packing per unit volume. In this experiment, the above parameters of “FCP-08” foam ceramic corrugated board packing were measured by Autosorb-3B nitrogen adsorbent instrument using the method of Brunauer–Emmett–Teller (BET) nitrogen adsorption capacity. The main structural parameters are shown in Table 2.

Table 2.

FCP-08 foam ceramic corrugated board packing block’s structural parameters.

Structural parameters	Value
Through-hole’s average pore size (mm)	45
Specific surface area $(m^{2} / m^{3})$	16,871
Effective porosity (%)	93
Packing block’s upwind size (length × width) (mm)	680 × 680
Filler blocks’ height (mm)	1000

By analyzing the data in the table, it can found that the value of specific surface area and porosity of “FCP-08” packing are relatively high. This is because the packing is made of foamed ceramic corrugated board in a crossed manner. Except for the formation of larger through-hole, the foam ceramic corrugated board itself has a special three-dimensional connected hole network structure, making the specific surface area and the porosity of the packing increased significantly.

Experimental devices and the processes

The thermal and mass transfer experiment device of the foam ceramic packing cooling tower established in this study is shown in Figure 3. It mainly includes thermostatic water tank with an electric heater (1), circulating water pump (2), electromagnetic flow meter (3), water distributor (4), cooling tower with “FCP-08” packing (5), collecting pool (6), centrifugal blower (7), air distribution chamber (8), air inlet heater (9), steam humidifier (10), air inlet rectifier grid (11), drift eliminator (12), thermostat (13), electric heater (14), bypass pipe (15), filter (16), temperature measuring device (17), differential pressure measuring device (18) (inclined tube pressure gauge), water tank feed inlet (19), and a number of valves. In order to improve the visualization degree of the experiment, the cooling tower body is assembled with transparent Plexiglass plate. The external structural size of the tower body is 700 mm × 700 mm × 2500 mm, and the size of the “FCP-08” packing inside the tower is 680 mm × 680 mm × 1000 mm. The cooling tower adopts tube pressure water distribute, the water distributor consists of five equally spaced DN50 steel tube, the bottom of each steel tube is equally spaced with an opening hole and equipped with a nozzle, and the diameter of the hole is 7 mm.

Figure 3.

Experimental apparatus for heat and mass transfer of ceramic packing cooling tower.

The work flow and principle of the experiment are as follows: when spray water in the constant temperature water tank is cyclically heated to the set spray temperature into the tower, it flows through the circulating pump and the flow meter into the top of the cooling tower. After measuring the inlet water temperature $(T_{i})$ of the cooling tower, the spray water evenly is sprayed on the experimental packing through the water distributor. During the process of downing through the packing, the hot water is cooled by contact with air and then falls into the collecting pool, and the spray water back flows to the constant temperature water tank for heating and recycling after measuring the outlet water temperature $(T_{o})$ of the cooling tower. Due to evaporation, the water level of the constant temperature water tank gradually decreases and the water is added to the tank through the water tank feed inlet. The air from centrifugal fan flows into the air distribution chamber with a rectifying grid through air inlet heater and steam humidifier to measure the dry and wet bulb temperatures ( $t_{ai}$ and $τ_{ai}$ ) of the air before entering the packing. The air enters evenly from the bottom of the cooling tower, making heat and mass transfer with the spray water during the process of up through the packing, and the air humidity increases. Before leaving the tower, the air passes through the drift eliminator to capture the droplets of wet air entrainment and returns it to the packing in order to reduce the water drift rate. Before the air is discharged into the atmosphere, the outlet dry and wet bulb temperatures ( $t_{ao}$ and $τ_{ao}$ ) of the air are measured.

As shown in Figure 3, the dry and wet bulb temperatures in the inlet and outlet of the cooling tower are measured by a digital wet and dry bulb thermometer, the inlet and outlet water temperatures of the cooling tower are measured by platinum resistance temperature sensor, and the sensor is installed in the inlet section of the spray water temperature measurement casing and the pipeline from the collecting pool to the constant temperature water tank; each temperature is measured by multiple PT100 sensor nodes and the measurement results are averaged; the flow rate of spray water is measured by electromagnetic flow meter. The parameters of the experimental test instrument are shown in Table 3.

Table 3.

Experimental test instrument parameters.

Equipment	Type	Measuring range	Error
Electromagnetic flow meter	PC-LDY	0.3–6.0 m³/h	±1%
Wet and dry bulb thermometer	DT-8896	−10°C to 60°C	±0.30°C
PT100 temperature sensor	WZP-230	0°C–200°C	±0.50°C
Hot-wire anemometer	TSI9565	0–25 m/s	±3%

By adjusting the fan frequency converter to change the inlet air flow rate, the wind speed of measuring points in the cross section of the outlet straight pipe of the cooling tower is measured by the hot-wire anemometer using the method of homolographic stationing;¹³ taking the arithmetic average of the wind speed of each measuring point as the average outlet wind speed, the air mass flow rate $G$ passing through the packing of the cooling tower can be derived from the following equation

G = A ρ_{a} v_{0}

(1)

In order to ensure the intuition and comparability of this experiment, the inlet air dry and wet bulb temperatures of the cooling tower should be stably unchanged. In the course of the experiment, the water–air mass flow ratio and inlet water temperature are independent variables. As for the water–air mass flow ratio, considering the convenience of adjusting, the change is mainly achieved by controlling the inlet air mass flow rate to be constant and adjusting the mass flow rate of the spray water. In practical engineering applications, it is generally believed that the inlet water temperature is the dependent variable and only varies with other independent variables such as the heat load of the condenser or the mass flow rate of the spray water. However, in this experiment, the heat load has been independent, and the mass flow rate of the spray water remains unchangeable under the same work condition, thus the inlet water temperature can be taken as independent variable. The experimental research consists of 21 tests obtained by modifying the water mass flow in seven levels and the inlet water temperature in three levels. The mass flow rate of the spray water takes 7.0, 8.0, 9.0, 10.0, 11.0, 12.0, and 13.0 m³/h, and the inlet water temperature of the cooling tower takes 32°C, 35°C, and 38°C.

Experimental uncertainty

Considering uncertainties of individual instruments, it is significant to study their effect on the final result. The relative uncertainty is expressed as follows¹⁴

\frac{u_{ζ}}{ζ} = \sqrt{{{(\frac{1}{ζ} \frac{\partial ζ}{\partial x_{1}} u_{1})}^{2} + {(\frac{1}{ζ} \frac{\partial ζ}{\partial x_{2}} u_{2})}^{2} + {(\frac{1}{ζ} \frac{\partial ζ}{\partial x_{3}} u_{3})}^{2} + \dots + {(\frac{1}{ζ} \frac{\partial ζ}{\partial x_{n}} u_{n})}^{2}}}

(2)

In this article, uncertainties involved in different instruments used in this experiment for measurements are evaluated as follows

\begin{matrix} \frac{u_{v_{0}}}{v_{0}} = 0.0125, \frac{u_{L}}{L} = 0.014, \frac{u_{T}}{T} = 0.00019 \\ \frac{u_{t_{a}}}{t_{a}} = 0.00015, \frac{u_{τ_{a}}}{τ_{a}} = 0.00016 \end{matrix}

(3)

Using equations (2) and (3), relative uncertainties involved in different performance parameters based on experimental results can be calculated as follows

\begin{matrix} \frac{u_{G}}{G} = 0.0123, \frac{u_{R}}{R} = 0.0023, \frac{u_{ω}}{ω} = 0.0126 \\ \frac{u_{e}}{e} = 0.0018, \frac{u_{Q_{w}}}{Qw} = 0.0153, \frac{u_{m_{ev}}}{m_{ev}} = 0.0018 \\ \frac{u_{KaV / L}}{KaV / L} = 0.00189 \end{matrix}

(4)

Heat and mass transfer phenomena in the cooling tower

Cooling characteristic coefficient

In the packing of the cooling tower, the flow of water is dominated by gravity, and the flow of air depends on the natural convection. While the gas and water are carrying on the sensible heat transfer, the evaporative cooling is also carried out by absorbing the latent heat of vaporization due to the evaporation of water. The enthalpy difference method proposed by Michael united the heat transfer formula with the difference of temperature and concentration as the driving force to a heat transfer formula with the difference of enthalpy as the driving force.¹⁵ Merkel introduced the Lewis relation $(Le = 1)$ into the equation and combined the heat and mass dissipation in the enthalpy, deriving the heat dissipation equation with the difference of enthalpy as the driving force

d Q_{w} = Ka (h_{sw} - h_{a}) dV

(5)

The heat dissipation of water can also be expressed by the difference of the water temperature between the inlet and outlet of the packing, modifying the Michael equation by combining the equation of Berman¹⁶ and taking the heat coefficient of evaporated water $k$ into the equation considering the loss of water caused by evaporation. The heat dissipation of water is expressed as equation (6)

d Q_{w} = \frac{1}{k} C_{pw} Ldt

(6)

According to the literature,¹⁶ the heat coefficient of evaporated water $k$ can be calculated by empirical formula (7)

k = 1 - \frac{C_{pw} T_{o}}{\bar{r}}

(7)

\bar{r} = 2500.8 - 2.343488 \times \frac{T_{i} + T_{o}}{2}

(8)

The heat dissipating capacity of water is equal to the heat absorbing capacity of air in the packing. Combine equation (5) with equation (6) to establish the heat balance equation

\frac{1}{k} C_{pw} L = Ka (h_{sw} - h_{a}) dV

(9)

Integrating on both sides of equation (9)

\frac{KaV}{L} = \frac{1}{k} \int_{T_{o}}^{T_{i}} \frac{C_{pw} dt}{h_{sw} - h_{a}}

(10)

According to Merkel’s¹⁵ method, $K α V / L$ is the cooling characteristic coefficient. The integration of equation (10) is solved numerically to obtain the values of the cooling characteristic coefficient, $K α V / L$ , reported to the volume of the exchange core, for different experimental operating conditions for water and air. Since the temperature of water is not a direct function of the air enthalpy, it is difficult to integral directly. Here, Chebyshev integration method recommended by the American Cooling Technology Institute (CTI) is used to solve the problem, and the formula of cooling characteristic coefficient is (as mentioned in Perry et al.¹⁷)

\begin{matrix} \frac{KaV}{L} = \int_{T_{o}}^{T_{i}} \frac{C_{pw} dt}{h_{sw} - h_{a}} \\ = \frac{C_{pw} R}{4} (\frac{1}{Δ h_{1}} + \frac{1}{Δ h_{2}} + \frac{1}{Δ h_{3}} + \frac{1}{Δ h_{4}}) \end{matrix}

(11)

R = T_{i} - T_{o}

(12)

Among them, R refers to the temperature difference between the inlet water and outlet water of the cooling tower, which is known as the range of cooling water, where

\begin{matrix} Δ h_{1} = value of [h_{sw} at {T_{i} + 0.1 (T_{o} - T_{i})} - {h_{a, 1} + 0.1 (h_{a, 2} - h_{a, 1})}] \\ Δ h_{2} = value of [h_{sw} at {T_{i} + 0.4 (T_{o} - T_{i})} - {h_{a, 1} + 0.4 (h_{a, 2} - h_{a, 1})}] \\ Δ h_{3} = value of [h_{sw} at {T_{o} - 0.4 (T_{o} - T_{i})} - {h_{a, 2} - 0.4 (h_{a, 2} - h_{a, 1})}] \\ Δ h_{4} = value of [h_{sw} at {T_{o} - 0.1 (T_{o} - T_{i})} - {h_{a, 2} - 0.1 (h_{a, 2} - h_{a, 1})}] \end{matrix}

Cooling tower efficiency

The cooling efficiency of the counterflow wet cooling tower is usually evaluated by the heat exchange efficiency $e$ which can measure the degree of evaporative cooling in the cooling tower. Heat exchange efficiency is determined as the ratio of actual to maximum water temperature decrease

e = \frac{R}{T_{i} - τ_{ai}} \times 100 %

(13)

In formula (13), the inlet temperature of the spray water $T_{i}$ and the inlet wet bulb temperature of the air $τ_{ai}$ are the main factors affecting the cooling efficiency of the cooling tower.

Heat rejected

Considering convective and evaporative heat transfer along with water loss caused by evaporation, Kloppers¹⁸ proposed equation (14) (improved Merkel equation) to calculate the total heat rejected from water

Q_{w} = L C_{pw} T_{o} - (L - m_{ev}) C_{pw} T_{i}

(14)

where $m_{ev}$ refers to the mass flow rate of the evaporated water and can be determined using equation (15)

m_{ev} = G (ω_{o} - ω_{i})

(15)

Experimental results and analysis

Cooling water range and cooling tower efficiency

Figure 4 shows the variation in the cooling water range, $R$ , with the water–air mass flow ratio, $L / G$ , for inlet water temperatures of 32°C, 35°C, and 38°C. It is apparent from the figures that the cooling water range decreases gradually with an increase in $L / G$ for all inlet water temperatures. The reason for this change trend can be included as the increase in $L / G$ enables the per unit mass of inlet air stream to contact with more volume of spray water inside the cooling tower and the increase in the heat exchange resistance on the water side weakens the effect of contact heat transfer and the evaporation heat transfer at the same time. It can also be shown through these figures that the cooling water range increases with the increase in the inlet water temperature under the same water–air mass flow ratio and the decrease becomes more obvious at higher inlet water temperature than at lower inlet water temperature, which can be attributed to the value of saturation vapor pressure increase along with the increase in the water temperature; the evaporation rate of water is enhanced, and the heat transfer between the water and air is more intense. When the values of $L / G$ are sufficiently high, the mass flow rate of the sprayed water is large, the membrane flowing through the surface of the inclined through-hole is too thick, some water flows into the collecting pool directly not having enough time going through the three-dimensional connected hole in the packing, and the cooling time and the cooling area are reduced.

Figure 4.

The variation in cooling water range with the water–air mass flow ratio.

Figure 5 shows the variation in the cooling tower efficiency, $e$ , with the water–air mass flow ratio, $L / G$ , for inlet water temperatures of 32°C, 35°C, and 38°C. It is observed that the cooling tower efficiency decreases with increasing water–air mass flow ratio for all inlet water temperatures. The efficiency is higher for the lowest values of $L / G$ , which in fact refers to the best cooling of the circulating water. However, under similar conditions of ambient air, and operating parameters, higher efficiencies are obtained at higher inlet water temperatures for the same water–air mass flow ratio. Such evolution can be attributed to the efficiency formula (13); it can be seen that the lower the outlet temperature, the higher the efficiency of the cooling tower when the inlet water temperature and the inlet air wet bulb temperature are certain, that is, for the same inlet conditions, the greater the range of the cooling water, R, the higher the efficiency of cooling tower. And according to the analysis of the variation rule of the cooling water range before, the cooling water range decreases with the increase in the water–air mass flow ratio, and the efficiency of cooling tower also decreases.

Figure 5.

The variation in cooling tower efficiency with the water–air mass flow ratio.

Cooling characteristic coefficient

Figure 6 shows the variation in the cooling characteristic coefficient, $K α V / L$ , with the water–air mass flow ratio, $L / G$ , for inlet water temperatures of 32°C, 35°C, and 38°C, respectively. The cooling characteristic coefficient decreases with an increase in $L / G$ , whereas it increases with increasing inlet water temperature. Meanwhile, this decrease is less pronounced as $L / G$ increases. EI-Dessouky⁴ attributed the reason to the decrease in water evaporation of per unit inlet water. The volume of air cooling the unit mass water decreases with the increment of the mass flow rate of the spray water. On one hand, the thickness of the water membrane on the surface of the inclined through-hole increases and the resistance of the air flowing through the through-hole increases, thus the blockage of the through-hole is deteriorated and the heat transfer tends to be uneven. On the other hand, the flow velocity of continuous flow in the connected hole of the packing is increased and the heat transfer time is shortened. Combining the above two reasons, the evaporation rate of water is decreased, the heat transfer capacity is weakened, and the cooling characteristic coefficient is decreased.

Figure 6.

The variation in cooling characteristic coefficient with the water–air mass flow ratio.

As already mentioned, the cooling characteristic coefficient, $K α V / L$ , is influenced by the air and water flow rates. Therefore, in order to derive an equation characterizing the heat exchange through the “FCP-08”-type packing, the relationship between the cooling characteristic coefficient and the water–air mass flow ratio is fitted into the form of a power function for each inlet water temperature $(T_{i})$ , which is

\frac{K α V}{L} = c {(\frac{L}{G})}^{n}

(16)

$T_{i} = 32 \circ C$ : $K α V / L = 1.8898 (L / G)^{- 0.2129}$ , within a standard deviation of 0.84%.

$T_{i} = 35 \circ C$ : $K α V / L = 1.6940 (L / G)^{- 0.2316}$ , within a standard deviation of 1.21%.

$T_{i} = 38 \circ C$ : $K α V / L = 1.5083 (L / G)^{- 0.2269}$ , within a standard deviation of 1.04%.

The results were then compared with those obtained in the literature for other types of packing. These correlations are as follows:

The correlation given in Armfield Ltd¹⁹

\frac{K α V}{L} = 1.700 {(\frac{L}{G})}^{- 0.15}

where $L / G = 0.35 - 2.5$ .

Barile et al.’s²⁰ correlations:

Diameter of the particles, d = 19.05 mm

\frac{K α V}{L} = 1.118 {(\frac{L}{G})}^{- 0.514}

Diameter of the particles, d = 38.10 mm

\frac{K α V}{L} = 1.147 {(\frac{L}{G})}^{- 0.430}

where $L / G = 0.7 - 3$ .

Kelly and Swenson’s²¹ correlation

\frac{K α V}{L} = 0.07 + 1.62 {(\frac{L}{G})}^{- 0.62}

where $L / G = 0.6 - 3.5$ .

Gharagheizi et al.’s¹ correlations:

VCP

\frac{K α V}{L} = 0.342 {(\frac{L}{G})}^{- 0.253}

HCP

\frac{K α V}{L} = 0.224 {(\frac{L}{G})}^{- 0.295}

where $L / G = 0.2 - 4$ .

Commercial packing with S wave shape

\frac{K α V}{L} = 0.904 {(\frac{L}{G})}^{- 0.357}

where $L / G = 0.5 - 3.5$ .

The comparison results of the experimental results under the work condition of this article (inlet water temperatures 32°C, 35°C, and 38°C) with the experimental results of other researchers and commercial packing are shown in Figures 7 –9. It can be seen that the cooling characteristic coefficient in this research is larger than that of the HCP and VCP, simultaneously, lower than the value in Instruction manual.¹⁹Figure 8 shows that the cooling characteristic in this research is larger than that of pellicular regime (PR) and bubble and dispersion regime (BDR).²² In comparison with Kelly and Swenson,²¹ it can be seen that when the water–air mass flow ratio is larger than 1, the cooling characteristic coefficient in this research becomes larger than Kelly et al.’s. Figure 9 shows that the cooling characteristic in this research is larger than that of Barile et al.’s²⁰ and commercial PVC packing with S wave shape. The gap becomes larger with the increase in water–air mass flow ratio. In addition to the difference of packing structure, the height of packing in the experiment, the inlet water temperature, and the weather conditions are also the factors that cause the difference of the cooling characteristic coefficient.

Figure 7.

Comparison of cooling characteristic coefficient at 32°C.

Figure 8.

Comparison of cooling characteristic coefficient at 35°C.

Figure 9.

Comparison of cooling characteristic coefficient at 38°C.

This indicates clearly that the heat and mass transfer coefficients of the packed cooling tower used in this investigation are much higher than other types. The “FCP-08”-type packing has better water cooling capacity. The reason can be attributed to the three-dimensional connected hole structure which connects the two sides’ surface of corrugated plate very well and enhances the horizontal mixing and even distribution of the liquid phase fluid in the packing. Therefore, the cooling performance of the packing is improved. This suggests that cooling tower filled with the “FCP-08”-type packing possesses better heat and mass transfer characteristics and can contribute significantly to energy saving.

Heat rejected

Figure 10 shows the variation in heat rejected by the cooling tower, $Q_{w}$ , with the water mass flow rate, $L$ , carried out at different inlet water temperatures: 32°C, 35°C, and 38°C. It appears from these figures that the heat rejected by the cooling tower filled with the “FCP-08”-type packing increases progressively with increasing $L$ while $G$ is kept constant for all inlet water temperatures. The reasons for this rule can be explained partly by the increase in the evaporation of water per unit mass of inlet air stream. When the water mass flow ratio continues to increase, however, the increasing upward trend of the total heat rejected is obviously slowed down and is in large part due to the gradual decrease in cooling water range. In addition, it can be observed that the heat rejected by the cooling tower increases along with the increase in inlet water temperature under the same water mass flow ratio. Such evolution can be attributed to the energy transferred from hot water to cooling air especially at high water temperature. From the perspective of heat and mass transfer, the diffusivity rate of water evaporation increases with the increase in the inlet water temperature and the evaporation of water is strengthened, so that the heat and mass transfer is more intense and the total heat rejected increases.

Figure 10.

The variation in heat rejected by the cooling tower with water–air mass flow ratio.

Conclusion

In this study, evaporative heat and mass transfer inside a counterflow wet cooling tower filled with “FCP-08” foam ceramic packing has been investigated for a wide range of water–air mass flow ratio. The results obtained at different inlet water temperatures, 32°C, 35°C, and 38°C, can be summarized as follows:

The cooling water range, $R$ , and the cooling tower efficiency, $e$ , decrease with the increase in water–air mass flow ratio, $L / G$ , and increases with the increase in inlet water temperature. Higher values of $R$ and $e$ are obtained at lower values of $L / G$ . Indeed, the best water cooling is obtained at high inlet water temperature, lower water mass flow rate, or upper air mass flow rate.

This article considers the mass change in water caused by evaporation based on the theory of Merkel and selects the calculation method of Bierman. The cooling characteristic coefficient, $K α V / L$ , decreases with the increase in water–air mass flow ratio, $L / G$ , and increases with the increase in inlet water temperature. Meanwhile, the trend of decrease is less pronounced as $L / G$ increases and the value of $K α V / L$ is obviously higher than that of other packings investigated before.

The heat rejected by the cooling tower, $Q_{w}$ , increases with the increase in the water mass flow rate, $L$ , for different inlet water temperatures, and the increasing upward trend of total heat rejected is obviously slowed down as the water mass flow rate continues to increase. Thus, it can be concluded that the cooling tower rejects higher rates of heat from process cooling water at higher water temperature and mass flow rates.

Cooling tower filled with the “FCP-08”-type packing compared to systems filled with other types of packing leads to a better performance of heat and mass transfer characteristics and therefore a great water cooling capacity. This can contribute significantly to energy saving and economics.

Footnotes

Appendix 1

Handling Editor: Assunta Andreozzi

Declaration of conflicting interests

The author(s) declared no potential conflicts of interest with respect to the research, authorship, and/or publication of this article.

Funding

The author(s) disclosed receipt of the following financial support for the research, authorship, and/or publication of this article: This research was supported by the National Natural Science Foundation of China (key project; no. 51509076) and the Natural Science Foundation of Jiangsu Province (key project; no. BK20150816).

ORCID iD

Qing Jie Kong

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Experimental investigation of the heat and mass transfer phenomena in a counterflow wet cooling tower with foam ceramic packing

Abstract

Keywords

Introduction

Experimental materials and devices

Packing materials and structure

Packing materials

Packing structure

Experimental devices and the processes

Experimental uncertainty

Heat and mass transfer phenomena in the cooling tower

Cooling characteristic coefficient

Cooling tower efficiency

Heat rejected

Experimental results and analysis

Cooling water range and cooling tower efficiency

Cooling characteristic coefficient

Heat rejected

Conclusion

Footnotes

Appendix 1

Declaration of conflicting interests

Funding

ORCID iD

References