Reducing water consumption of an industrial plant cooling unit using hybrid cooling tower

Energy Conversion and Management 51 (2010) 311–319

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Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman

Reducing water consumption of an industrial plant cooling unit using hybrid cooling tower Ebrahim Rezaei a,*, Sirous Shafiei a, Aydin Abdollahnezhad b a b

Faculty of Chemical Engineering, Sahand University of Technology, P.O. Box 51335/1996, Tabriz, Iran Tabriz Refinery R&D, Tabriz, Iran

a r t i c l e

i n f o

Article history: Received 6 June 2008 Received in revised form 26 February 2009 Accepted 30 September 2009 Available online 5 November 2009 Keywords: Hybrid cooling tower Modeling Simulation Water loss reduction

a b s t r a c t Water consumption is an important problem in dry zones and poor water supply areas. For these areas use of a combination of wet and dry cooling towers (hybrid cooling) has been suggested in order to reduce water consumption. In this work, wet and dry sections of a hybrid cooling tower for the estimation of water loss was modeled. A computer code was also written to simulate such hybrid cooling tower. To test the result of this simulation, a pilot hybrid tower containing a wet tower and 12 compact air cooled heat exchangers was designed and constructed. Pilot data were compared with simulation data and a correction factor was added to the simulation. Ensuring that the simulation represents the actual data, it was applied to a real industrial case and the effect of using a dry tower on water loss reduction of this plant cooling unit was investigated. Finally feasibility study was carried out to choose the best operating conditions for the hybrid cooling tower configuration proposed for this cooling unit. Ó 2009 Elsevier Ltd. All rights reserved.

1. Introduction There are two types of cooling towers commonly used in various plants: wet and dry cooling towers. In general, cooling towers may operate either counter currently or with the air entering at the side of the tower and flowing across the water stream. The motive force needed for the air to pass the tower is provided naturally or by forced or induced draft fans. Wet cooling towers have been introduced as one of the direct contact heat exchangers based on the evaporation of water into air in a packing media. Although the water stream cools somewhat as a result of sensible heat, most of the heat rejection in a wet cooling tower is caused by evaporation of some water into the air [1]. One of the advantages of wet cooling towers over the dry towers is evaporative cooling where the water temperature may approach the atmospheric wet bulb temperature rather than the dry bulb temperature. However, use of wet cooling towers causes water loss which is an intolerable issue in dry regions with poor water supplies. Also visible plume may occur when the saturated exhaust air from wet cooling towers is confronted with ambient air in the case that the ambient air is very humid or the temperature is

* Corresponding author. Tel.: +1 306 2907418. E-mail addresses: [email protected] (E. Rezaei), shafi[email protected] (S. Shafiei). 0196-8904/$ - see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.enconman.2009.09.027

low. Nowadays, the visible plume attracts the public attention with increasing concern of environmental issues. For example under the air pollution control ordinance of Hong Kong, the visible plume is not allowed [2]. The visible plume can be extremely serious extending up to 100 m and some times causes visibility and darkness under unfavorable atmospheric conditions. It is also a nuisance and may result in misconception of a fire accident or hazardous pollutant for the local people in the highly dense cities [2]. Dry cooling towers include compact heat exchangers in which water flows in tubes and air passes through the exchangers. These towers have no water loss and are very suitable to operate under cold and moderate climates. Except of their high investment cost, the main problem with dry towers is that their efficiency decreases with the increase of the ambient air temperature. This occurs in hot summer days when the heat transfer driving force decreases. Much attention has been paid to issues on wet cooling towers relating to modeling and simulation of wet cooling towers [3–6], study the effect of different packing types on the characteristics and thermal performance of cooling towers [7–8], use of neural network in the prediction of cooling tower performance [1], evaluation of thermal efficiency of cooling towers with respect to fouling [9] and other aspects which are mostly related to the design of new wet cooling units. However, little attention has been placed on revamping of existing wet cooling units and finding solutions for their conventional

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Nomenclature Wet section specific heat, J/(kg K) Cp heat transfer coefficient, W/(m2 K) hc mass transfer coefficient, kg/(m2 s) hd i enthalpy, J/kg enthalpy of air–water mixture at ambient air temperaima ture, J/kg enthalpy of saturated air at bulk water temperature, J/kg imasw enthalpy of supersaturated air, J/kg iss volumetric mass transfer coefficient, kg/(m3 s) Kya Lewis umber, h/(cp hd) Lef m mass flow rate, kg/s NTU number of transfer units T temperature, °C or K V volume of fill, m3 w humidity, kg water vapor/kg dry air Subscripts a air i inlet o outlet v vapor w water Dry section bare area, m2 Ab Af finned area, m2 Afr frontal area, m2 Amin minimum free flow area, m2 At total heat transfer area, m2 Au tube area between fins, m2 Cc heat capacity flow rate of a cold stream, W/K heat capacity flow rate of a hot stream, W/K Ch specific heat, J/(kg K) Cp inside diameter of tubes, m di equivalent diameters of fins, m df outside diameter of tubes, m dr F correction factor G mass velocity, kg/(m2 s) fin height, m Hf water side heat transfer coefficient, W/(m2 K) hi air side mean heat transfer coefficient, W/(m2 K) ho

problems which are generation of plume and loss of water. According to decrease in natural water sources and severity of laws on plume generation in recent years, the best choice to improve the economics of wet cooling towers and their operability is to add a dry cooling tower to existing wet cooling units called a hybrid cooling tower. This paper points this issue and suggests a procedure to achieve the above objectives using a hybrid cooling tower. It is organized in the following manner: modeling and simulation of wet and dry sections are first presented in Sections 2 and 3. Section 4 describes pilot construction and its characteristics while experiments conducted on this pilot are explained in Section 5. Section 6 shows the application of the procedure to an industrial case by addition of air cooled heat exchangers to Tabriz Refinery cooling unit and estimating the amount of water reduction in various configurations. Section 7 summarizes the results and gives the most economic option for design of the plant hybrid cooling tower.

ho,b

ho,t k L N Nu nr nt Pr p Q Re r re s T Ufr Umax Uo,b Uw w

effective air side heat transfer coefficient based on bare area and with respect to fin efficiency, W/(m2 K) effective air side heat transfer coefficient based on total area and with respect to fin efficiency, W/(m2 K) thermal conductivity, W/(m K) length of each tube, m total number of tubes Nusselt number number of rows number of tubes in each row Prandtl number tube pitch, m heat load, W Reynolds number outside radius of tubes, m equivalent radius of fins, m fin spacing, m temperature, °C or K inlet air speed, m/s maximum air speed at minimum surface area, m/s overall heat transfer coefficient based outside bare area, W/(m2 K) water velocity in tubes width of fins, m

Greek symbols fin efficiency viscosity, Pa s density, kg/m3 Amin/Afr DP pressure drop, Pa DT effective temperature difference, °C DTlm logarithmic mean temperature difference, °C

gf l q r

Subscripts a air i inlet o outlet w water

2. Wet section modeling and simulation The wet section of a hybrid cooling tower is a wet cooling tower that cools water by the combination of heat and mass transfer. The water to be cooled is pumped into the tower and is distributed into a fill in which air is passed either naturally or mechanically. The Poppe method accurately predicts the water content of the exit air [10–11] which is a very important factor in the design of hybrid cooling towers [12]. This method leads to the following four equations for unsaturated and supersaturated air [3]:

C pw mmwa ðwsw  wÞ dw ¼ dT w imasw  ima þ ðLef  1Þ½imasw  ima  ðwsw  wÞiv   ðwsw  wÞC pw T w ð1Þ

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  dima mw C pw ðwsw  wÞC pw T w ¼ 1þ dT w ma imasw  ima þ ðLef  1Þ½imasw  ima  ðwsw  wÞiv   ðwsw  wÞC pw T w

ð2Þ

C pw mmwa ðwsw  wsa Þ dw ¼ dT w imasw  iss þ ðLef  1Þ½imasw  iss  ðwsw  wsa Þiv þ ðw  wsa ÞC pw T w  þ ðw  wsw ÞC pw T w

ð3Þ

  dima mw C pw T w ðwsw  wsa Þ ¼ C pw 1þ imasw  iss þ ðLef  1Þ½imasw  iss  ðwsw  wsa Þiv þ ðw  wsa ÞC pw T w  þ ðw  wsw ÞC pw T w dT w ma

ð4Þ

Eqs. (1) and (2) represent the differential changes of humidity and enthalpy of the unsaturated air as a function of water temperature through the tower while Eqs. (3) and (4) express differential changes of the supersaturated air. The Lewis numbers in these equations are calculated by Eqs. (5) and (6) for the unsaturated and supersaturated air respectively [3]:

Lef ¼ 0:86550:667

Lef ¼ 0:86550:667

    wsw þ 0:662 wsw þ 0:662 1 Ln w þ 0:662 w þ 0:662

ð5Þ

    wsw þ 0:662 wsw þ 0:662 1 Ln wsa þ 0:662 wsa þ 0:662

ð6Þ

Inlet air temperature (°C) Inlet air humidity (kg/kg) Inlet water temperature (°C) Outlet water temperature (°C) Air flow rate (kg/s) Water flow rate (kg/s)

25.0 0.0100 40.0 21.1 650.0 450.0

Table 2 Comparison of the results.

The fourth order Runge–Kutta method was employed to solve Eqs. (1) and (2) or (3) and (4). The first step in the solution process is to divide the fill into intervals where the water temperature difference is equal across each [3]. For example, five intervals are considered in Fig. 1 for the simulation. To start solving Eqs. (1) and (2) or (3) and (4), a value for the exit air humidity (wo) is given as an initial guess and a new value is subsequently determined by solving Eqs. (1) and (2) or (3) and (4). This procedure continues until the calculated exit air humidity converges. Results show that if suitable estimation is used, convergence will be reached after two or three iterations [3]. In each interval the state of air should be known (unsaturated or supersaturated) in order to select proper Eqs. (1) and (2) or (3) and (4). It is obvious that if the air becomes supersaturated at one of the intervals, it will remain in the supersaturated state thorough the rest of the fill. Results of the simulation were compared with a previous work [13] in order to validate the method used in this work. Table 1

Fig. 1. Counter flow fill divided into five intervals.

Table 1 Input of this work and the previous work [13].

Outlet air temperature (°C) Outlet air humidity (kg/kg) Evaporation (kg/s)

Current simulation

Previous work [13]

31.5 0.030 12.5

30.7 0.029 12.4

shows the input of this simulation while Table 2 summarizes the output. A good agreement can be seen by the comparison of this work with [13] in the estimation of outlet conditions. It is good to note that although profiles of the air temperature and humidity are available by the simulation, only the exit air condition is presented for the comparison.

3. Dry section modeling and simulation The dry section of a hybrid cooling tower includes compact heat exchangers which are widely used in the industry especially as gas-to-gas or liquid-to-gas heat exchangers. Some examples are vehicular heat exchangers, condensers and evaporators in air conditioning, oil coolers, radiators, intercoolers of compressors and aircraft and space applications [14]. In such exchangers, the heat transfer surface area is increased by fins to increase the area per unit volume of an exchanger. Compact heat exchangers are divided into plate-fin and tube-fin heat exchangers. In the plate-fin type, each channel is defined by two parallel plates separated by fins or spacers. Fins or spacers are sandwiched between parallel plates. Fins are attached to the plates by brazing, soldering or welding [14]. In a tube-fin exchanger, round, rectangular and elliptical tubes are used and fins are employed either on the outside or on the inside, or on both sides and inside of the tubes, depending on the application. In a gas-to-liquid exchanger, which is employed in this work, the gas side heat transfer coefficient is very low compared with the liquid side heat transfer coefficient; therefore no fins are needed on the liquid side [14]. Fins on the outside tubes may be categorized as (1) flat or continuous (plate, wavy or interrupted) external fins on an array of tubes, (2) normal fins on individual tube, and longitudinal fins on individual tube [14]. The first step in the design of an air cooled heat exchanger is determination of the heat exchanger geometry. For continuous

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plate-fin compact exchangers with the tube layout angle (TLA) of 30° (Fig. 2), an equivalent diameter is given by [14]: 1 re ¼ 1:27 uðb  0:3Þ2 r

ð7Þ

where

p1 2r  1 1 p2 2 p22 þ 1 b¼ p1 4

u¼

ð8Þ ð9Þ

In the above equations, re and r are equivalent radius of fin and outside radius of tubes respectively. Having re, one can use the correlation of circular fins for the calculation of the total area of finned tubes [14]:

At ¼ A f þ Au ¼



 N pL 1 2 2 df  dr þ df w þ dr s ðs þ wÞ 2

ð10Þ

Table 3 shows the geometry of one of twelve heat exchangers used in the dry section of the pilot. Air side mean heat transfer coefficient (ho) can be calculated by the Colburn j-factor which is defined as a function of the Reynolds number [14,15]:

ho ¼

Re ¼

j C p Gmax

ð11Þ

2=3

Pr

qa U max dr qa dr U fr U fr Afr qa dr ¼ ¼ la la r Amin la

ð13Þ

Amin is defined by [14]:

In which:

j ¼ f ðJ p Þ;

Fig. 3. Determination of j by Jp.

J p ¼ Re

0:4



At Ab

0:15 ð12Þ

This functionality is indicated in Fig. 3 in which j-factor is plotted versus Jp for various fin spacing. In Eq. (12), the Reynolds number is determined based on the maximum air speed in the exchangers:



2wHf Amin ¼ nt L P1  dr  wþs

ð14Þ

Effective air side heat transfer coefficient based on total area and bare area of tubes is expressed by [14]:

ho;t ¼ ho



ho;b ¼ ho;t

gf Af þ Au

 ð15Þ

At At Ab

ð16Þ

gf is given by:

gf ¼

Tanhðmr/Þ mr/

ð17Þ

where

dr r¼ ; 2

r i r h e e ; /¼  1 1 þ 0:35Ln r r

sffiffiffiffiffiffiffiffiffi 2ho m¼ wkf

ð18Þ

To find water side heat transfer coefficient (hi), Dittus–Boelter correlation is used for turbulent flow in tubes [16]:

hi di ¼ Nu ¼ 0:023 Re0:8 Pr0:3 ; kw Fig. 2. Tubes with rotated triangular configuration.

qw di U w lw

ð19Þ

Neglecting fouling, thermal resistance and contact thermal resistance between fins and tubes, the overall heat transfer coefficient based on outside bare area of tubes can be calculated by the following equation [14]:

Table 3 Geometry of exchangers. Parameter

Value

Parameter

Value

Parameter

Value

Ab Af Afr Amin At Au di df

0.9576 m2 1.9626 m2 0.18 m2 0.0992 m2 2.7264 m2 0.7638 m2 0.0102 m 0.0163 m

dr Hf L N nr nt P1 P2

0.0127 m 0.179 m 1m 24 4 6 0.03 m 0.027 m

P3 r re s w

0.033 m 0.635 m 0.814 m 0.0018 m 0.0005 m 0.55

r

Re ¼

    1 1 dr dr dr 1 ¼ Ln þ þ U r;b hi di ho;b 2ktube di

ð20Þ

Having Uo,b, the heat load of the exchanger is defined as:

Q ¼ U r;b Ab DT m

ð21Þ

The effective temperature difference (DTm) can be obtained by Fig. 4 for which NTU, R and h are [17]:

E. Rezaei et al. / Energy Conversion and Management 51 (2010) 311–319

315

Fig. 4. Determination of the effective temperature difference (DTm).

NTU ¼

AU ; C min

R¼

Cc ; Ch

h¼

DT m ; T w;i  T a;i

C min ¼ minfC c ; C h g

1

qa ð22Þ

For flow normal to finned tube banks, total pressure drop of air, namely, the difference between the pressures at the inlet and outlet is given by [18]:

DP ¼ ðK a þ nr K f Þ

  1 qa U 2max 2

K f ¼ 4:71 Re0:286

þ

qa;o

! ð26Þ

According to the above equations, a computer code was written for the simulation of the dry section. This code is able to calculate the heat transfer area and air side pressure drop for a specified cooling duty.

4. Pilot set-up

ð24Þ  0:51  0:536  0:36 Hf P1  dr dr s P2  dr P1  dr

qa;i

1

ð23Þ

In which Ka and Kf are defined by:

Ka ¼ 1 þ r 2

¼ 0:5

1

ð25Þ

In Eq. (23), air density is evaluated at the average temperature between the inlet and outlet or it can be averaged between the inlet and outlet as [14]:

Based on the governing equations of Sections 2 and 3, a pilot (Fig. 5) was designed and constructed. Dry section of this pilot includes 12 compact air cooled exchangers. These exchangers can be used in a parallel or series arrangement with the wet section. Tubes have triangular configuration and they are made of aluminum. Eleven copper plate-fins are installed per inch of tubes with the dimension of 11.6  18 cm2. Wet section contains five packing decks which has 1 m height with the cross sectional area of

Fig. 5. Pilot view.

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0.66 m2. Inlet water is pumped to the top of the tower and is distributed by a distributor to pass the packing media. Fig. 6 shows the schematic diagram of this pilot with the following features:

Table 4 Characteristics of the instrumentation.

1. Inlet air. 1.1 Inlet air to the dry section. 1.2 Inlet air to the wet section. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 15. 16. 17. 18.

Inlet water to the dry section. Outlet water from the dry section. Inlet water to the wet section. Outlet water from the wet section. Compact heat exchangers. Outlet air from the pilot fan. Pump. Fan. Water distributors in the wet section. Packing media. Flexible hoses used to achieve different configurations in the dry section. Indirect steam heater. Water flow meter. Ball valve. Check valve. Collecting pool. Steam line.

To provide hot water for the system, a storage tank and an indirect steam heater are employed to heat the water to desired temperatures. The hot water is drawn from the storage tank by means of a pump and is routed to the cooling tower. A check valve is mounted after the pump to prevent possible damages. Two ball valves are placed in the inlet of dry and wet towers to direct water to each section. Therefore it is possible to either analyze the performance of dry and wet sections separately or simultaneously. Also these valves can be used to control the flow rate of the entering water to the pilot. A fan is used to take air into the tower to supply the coolant fluid of the pilot. Air enters from the bottom (see 1–2 in Fig. 6) of the pilot to the wet section and from each side of the system (see 1–1 in Fig. 6) to the dry section. The fan can provide different flow rates by setting its blades angle in the range of 15–45°. For collecting the cooled water from the wet cooling section of the pilot, a little pool was put at the bottom of the pilot.

Measured variable

Instrument

Accuracy

Air velocity (m/s) Water temperature (°C) Air temperature (°C) Air relative humidity (%) Water flow rate (mL/min)

Portable digital anemometer Digital thermometer Digital thermometer Portable digital hygrometer Digital turbine flow meter

±0.01 ±0.1 ±0.1 ±0.1 ±40

Table 5 Pilot dynamic characteristics. Inlet air flow rate into the exchangers (m3/s) Inlet air flow rate into the wet tower (m3/s) Total inlet air flow rate into the system (m3/s) Total pressure drop (Pa) System power (W) Fan power (W) Fan efficiency (%) Measured air side pressure drop in exchangers (Pa) Estimated air side pressure drop in exchangers by dry section simulation (Pa) Measured air pressure drop in the wet tower (Pa)

10

1-1

Three sets of experiments are explained in this section to examine the accuracy of wet and dry section simulations and to determine the characteristic of the filled media of the wet tower.

6 1-1

11

15 T

15

3

2 Hot Water

12 1-2

1-2

T 17 T 5

126.69

5. Experiments

4

6

4.00

This pilot has been equipped with several instruments in order to be able to accurately measure various parameters in experimental phase which will be described in Section 5. Table 4 summarizes the type and accuracy of measurement devices used in this pilot. To find the total air flow rate entering to the pilot, air velocity was measured at various points of the fan exhaust. Then an average value was calculated as the final value. The same procedure was used to measure the air flow rate to the dry section and the relative humidity of the discharged air from the fan. It is obvious that the air flow rate to the wet section is equal to the difference between the total and dry section air flow rates. Dynamic characteristics of this system are shown in Table 5 in which a good agreement is observed between the real air side pressure drop in compact exchangers and the pressure drop estimated by the simulation.

T 7 H 9

4.00 0.89 4.89 155 757.83 1200 63 28.32 21.2

14

16 8

Fig. 6. Schematic diagram of the experimental hybrid cooling tower.

13 18

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Sections 5.1 and 5.2 include two types of experiments which were designed to find out how much the models of Sections 2 and 3 are accurate. By knowing the errors of these models, two correction factors can be defined which enables the applicability of the current models to large and real scale cooling towers. The aim of Section 5.3 is to identify the volumetric mass transfer coefficient of the wet section and to obtain a correlation for expressing the number of transfer units (NTU) as a function of water and air mass flow rates. In order to be able to apply the model to an industrial case (Tabriz Refinery) the packing of the Tabriz Refinery cooling towers were used in wet section experiments. 5.1. Determination of dry section correction factor To validate the dry section model, experimental heat loads of exchangers were compared to the predicted values from the simulation. To have a variety of heat loads, different number of exchangers were put into service in each experiment. During experiments, inlet air velocity into the exchangers was constant and the number of exchangers was selected from 2 to 12 to work in series arrangements. Hot water was pumped to compact exchangers at different flow rates and relatively constant temperature and exited the pilot after transferring heat to the air. A correction factor is defined by dividing the experimental cooling load to the heat load obtained by the simulation. Table 6 shows the details of experiments which lead to an average correction factor of 1.12. In fact the error of modeling is 12% which is acceptable in engineering practice.

ter to the wet section. Results are indicated in Table 7 in which the average error of 6% is observed.

5.3. Wet section characterization According to the famous equation of Merkel [15,19–20], characterization of a wet cooling tower is given by the following equation which is equal to the number of transfer units (NTU):

NTU ¼

K y aV ¼ mw C pw

Z

dT w ðimasw  ima Þ

ð27Þ

This term (NTU) can be specified by plotting KyaV/(mwCpw) versus mw/ma for a constant inlet air flow rate [15]. The proposed relation in [15] is in the form of KyaV/(mwCpw) = a(mw/ma)b, in which a and b must be calculated from experimental data. It is important to know that the normal values of b are bounded in [1, 0.4]. It is notable that if the inlet air flow rate changes, the tolerance of this curve will be about ±20%. Also change in inlet water temperature of about ±5.5 °C does not affect the curve [15]. Five experiments were conducted to estimate KyaV/(mwCpw) from Eq. (27). Having NTU, one can determine coefficients a and b by using the least square method. In all experiments air flow rate was kept constant and variations of inlet water temperature were limited to less than 5.5 °C. Results are shown in Table 8 and the plot of KyaV/(mwCpw) versus mw/ma is available in Fig. 7, in which the wet section characterization is determined by KyaV/ (mwCpw) = 0.55(mw/ma)0.40. The number of wet towers which

5.2. Determination of wet section correction factor By dividing the exit air humidity of the wet tower to the humidity calculated by the wet section simulation, the second correction factor is determined. Three experiments were carried out to measure the air outlet humidity from the fan. In each experiment, humidity and temperature of inlet air and temperature of inlet water were constant. But different water flow rates were set to en-

Table 8 Experimental data of (KyaV/mwCpw). Inlet air temperature (°C) Inlet air relative humidity (%) Inlet water temperature (°C) Outlet water temperature (°C) Water flow rate (kg/s) Air flow rate (kg/s) mw/ma KyaV/(mwCpw)

28.0 10.0 48.5 23.0 0.116 1.04 0.111 1.3

28.0 21.0 48.0 25.0 0.144 1.04 0.138 1.2

29.0 10.0 50.0 26.0 0.212 1.04 0.204 1.1

28.0 9.4 48.0 28.0 0.303 1.04 0.292 0.9

31.0 6.7 47.0 29.0 0.381 1.04 0.367 0.8

Table 6 Determination of dry section correction factor. 2

5

Inlet water temperature (°C) Outlet water temperature (°C) Water flow rate (kg/s) System heat load (kW) Heat transfer area (m2) Inlet air temperature (°C) Water velocity in tubes (m/s) Air velocity at face area (m/s) Estimated heat load by simulation (kW) Correction factor Averaged correction factor

50.0 44.0 0.188 4.7 5.42 29.0 0.382 1.87 4.2 1.12

49.0 49.0 40.0 33.0 0.258 0.196 9.7 13.1 13.56 24.4 29.0 29.0 0.525 0.399 1.86 1.85 9.1 11.6 1.06 1.13 1.12

9

12 48.0 31.0 0.224 15.9 32.54 29.0 0.455 1.85 13.7 1.16

Table 7 Determination of wet section correction factor. Inlet air temperature (°C) Inlet air humidity (kg/kg) Inlet water temperature (°C) outlet water temperature (°C) Water flow rate (kg/s) Air flow rate (kg/s) Outlet air humidity (kg/kg) Estimated outlet air humidity (kg/kg) by simulation Correction factor Averaged correction factor

18.0 0.0026 49.0 22.0 0.028 0.08 0.016 0.017

18.0 0.0026 49.0 25.0 0.042 0.08 0.022 0.021

18.0 0.0026 49.0 28.0 0.056 0.08 0.030 0.025

0.94

1.04 1.06

1.20

1.4

1.2

(KyaV/MwCpw)

Number of exchangers

1

0.8

y = 0.55x

-0.40

2

R = 0.97

0.6 0.09

0.14

0.19

0.24

0.29

(Mw/Ma) Fig. 7. Wet tower characterization curve.

0.34

0.39

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should be used for a specified cooling duty in the refinery hybrid cooling tower can be calculated by this relation. 6. Application to an industrial case In this section, the models will be applied to an industrial case (Tabriz Refinery) in order to determine the following objectives:  The effect of addition of air cooled heat exchangers in parallel or series arrangements on the water consumption of the refinery wet cooling towers.  Required heat transfer area in each arrangement.  Operating conditions of wet and dry sections.  Economical analysis of the project. Tabriz Refinery was designed in 1974 and started on February 1977. This firm is placed in the south-west of Tabriz city, its nominal capacity is 110,000 bbl/day. It has a train of seven wet cooling towers to supply cooling water for the rejection of heat from hot process streams. Total water loss of the refinery wet towers is estimated to be about 304 m3/h in summer and 295 m3/h in winter. An inlet air temperature of 29 °C and 0.0043 of humidity is assumed in summer while 10 °C and 0.01 humidity is considered for air conditions in winter. It is good to note that to be in safe side, all estimations were carried out for hot days. Also total water loss consists of the evaporation of water plus entrained water [19]. 6.1. Hybrid ratio Hybrid ratio is defined as the ratio of the heat that must be removed in dry section to the total cooling load:

H:R: ¼ Q dry =Q total

ð28Þ

Increase of H.R. causes the water loss to be reduced but dictates use of more heat transfer area in the dry section. 6.2. Summer operation As mentioned before, inlet temperature of air and its humidity were assumed 29 °C and 0.0043 to simulate the cooling unit in summer. Results of the simulation for two different arrangements are presented in Tables 9 and 10. In parallel arrangement, each section cools the water to its final temperature (31 °C) separately but in the series arrangement, the water first enters into the dry section at 47 °C and is cooled due to the Hybrid ratio. It then goes to wet towers to be cooled to 31 °C. As Tables 9 and 10 show, parallel arrangement has lower water loss rather than the series one in each hybrid ratio. According to these tables, parallel arrangement needs larger heat transfer area

Table 9 Parallel arrangement (summer). Hybrid Ratio

0.2

0.4

0.6

0.8

Wet cooling load (MW) Total water loss (m3/h) Required area (m2)

147.691 243 98,231.8

110.768 182 194,666

73.846 121.8 294,703

36.923 61.2 392,936

0.2 147.691 252.6 51,161.7

0.4 110.768 199.2 102,339

0.6 73.846 144.7 173,950

0.8 36.923 89.6 276,274

Table 10 Series arrangement (summer). Hybrid ratio Wet cooling load (MW) Total water loss (m3/h) Required area (m2)

Table 11 Parallel arrangement (winter). Hybrid ratio Wet cooling load (MW) Total water loss (m3/h) Required area (m2)

0.2 147.691 213.8 22,510.4

0.4 110.768 171 45,022.1

0.6 73.846 114.1 67,539.9

0.8 36.923 57 89,996.8

1 0 0 112,556

Table 12 Series arrangement (winter). Hybrid ratio Wet cooling load (MW) Total water loss (m3/h) Required area (m2)

0.2 147.691 220.8 20,464.8

0.4 110.768 176.5 40,905.5

0.6 73.846 124.5 61,394.1

0.8 36.923 77 81,848.7

than series arrangement in different hybrid ratios. This is because of the poor heat transfer driving force in the dry section of parallel cases that causes the required area to be increased. 6.3. Winter operation Because the ambient air temperature is low enough, total heat can be removed in the air cooled exchangers in winter operation. Tables 11 and 12 include the results of the simulation in winter. Like the case of summer, the water loss for the parallel arrangement is less than the series one in each hybrid ratio. By comparison of the summer and winter results, it is obvious that the water loss of winter is less than the summer because of the decrease of ambient air temperature allowing the total heat to be rejected in the dry section. 6.4. Economic evaluation Investment and operating and maintenance costs of natural draft concrete cooling towers can be estimated by the following equations [21]:

Capital cos tð$Þ ¼ 1010 x3  105 x2 þ 70:552x þ 61609

ð29Þ

O&Mð$=yearsÞ ¼ 8  106 x2 þ 13:291x þ 13850

ð30Þ

In these equations x is the flow rate of water in gallon per minute (gpm). It must be mentioned that Eqs. (29) and (30) are proposed for towers with a temperature difference of 6 °C between inlet air and outlet water [21]. The average of this difference is 11.5 °C in this work which may cause a little overestimation in tower costs. Unit costs of 0.0396 $/m3 and 0.06 $/(kW h) were considered for cooling water and electricity in order to evaluate the reduction of wet towers operating cost. 7. Conclusion 1. The parallel arrangement in summer has a little thermal efficiency and requires large amounts of area for heat transfer. 2. In hybrid ratios equal to 0.2–0.4 series arrangement in summer is more suitable because of relatively high driving force. 3. In order to have a meaningful basis for analysis, the surface area of the dry cooling tower needed to cool water in winter without using wet towers of Tabriz Refinery was calculated to be 112556 m2. 4. With this area of 112,556 m2, in summer one of the following cases can be chosen: a. Series arrangement with a hybrid ratio equal to 0.43 and 37% water loss reduction in summer. In this case, three towers are required from seven towers of the refinery. b. Parallel arrangement with a hybrid ratio equal to 0.23 and 23% water loss reduction in summer. In this case, four towers are required from seven towers of the refinery.

E. Rezaei et al. / Energy Conversion and Management 51 (2010) 311–319

5. Economic analysis shows that if a dry tower with 112,556 m2 of heat transfer area is installed, the capital investment will be $1.3  106. For a rate of interest of 20%, parallel arrangement does not return this investment and for series arrangement a pay back period of 7.2 years is required. 6. If the price of fresh water is doubled with respect to the reduction of water sources and environmental impacts in the future, economic analysis shows that the pay back period will be 12.1 and 5.2 for parallel and series arrangements respectively. Therefore the installation of a dry tower with this area will be more flexible.

Acknowledgments Authors wish to thank Tabriz Refinery for their financial support for this research project. Also preliminary work done by Mr. A. Abbaspour, Ms. L. Mohammadi and Ms. S. Soltani for the design and construction of the pilot set-up is greatly appreciated. References [1] Hosoz M, Ertunc HM, Belgurcu H. Performance prediction of a cooling tower using artificial neural network. Energ Convers Manage 2007;48: 1349–59. [2] Xu X, Wang S, Ma Z. Evaluation of plume potential and plume abatement of evaporative cooling towers in a subtropical region. Appl Therm Eng 2008;28: 1471–84. [3] Kloppers JC, Kroger DG. A critical investigation into the heat and mass transfer analysis of counterflow wet-cooling towers. Int J Heat Mass Trans 2005;48: 765–77. [4] Naphon P. Study on the heat transfer characteristics of an evaporative cooling tower. Int J Heat Mass Trans 2005;32:1066–74. [5] Fisenko SP, Brin AA, Petruchik AI. Evaporative cooling of water in a mechanical draft cooling tower. Int J Heat Mass Trans 2004;47:165–77.

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