Each fan has only one design point which is established by a specific air flow, total pressure, air density, and fan speed. Starting with these data, it is possible to determine one planform and the twist distribution which will accomplish the required work with minimum horsepower.

To move air, the fan must overcome two resistances, which are measured as pressure drops across the fan. The first is a parasitic loss called the velocity pressure loss and this is the energy required to move the required air quantity without doing any work to overcome the system resistance. Work is done however to move the hot air away from the equipment.

The second resistance is the static pressure loss. It is the accumulated losses due to inlet louver, fill, drift eliminator, and fan inlet pressure drop, etc. This would be the work to be accomplished and reflects the design of the total system. Whether the air is distributed evenly across the fan is primarily a function of the blade and hub design. A properly designed blade will have adequate chord width and twist to ensure an even distribution of velocity pressure over its entire length. A properly designed hub will include a center air-seal disk which prevents negative air flow into the center of the fan. Following are two cases that most everyone has faced;

• Designing a new tower from sketch, attempting to get the best design possible.
• Replacing a fan on an old tower where practically nothing is known.

What we are seeking in the brand new selection procedure is an optimum fan diameter, number and type of blades, required pitch angle, fan rpm, and some estimate of horsepower. In some cases we are looking for an estimated sound-pressure level to satisfy EPA requirements for working area noise levels, more importantly noise level at a plant boundary, or a given noise sensitive location such as a residential area.

A problem that arises frequently is a fan replacement for an old tower that practically has no design data available. In this case, the only approach is to calculate the curve horsepower that will allow for the actual gear ratio and approximate density. We are looking for a fan that at least will be an adequate replacement for the original fan. Selecting a Hudson fan to replace an existing fan is generally very simple. The factors that must be known when replacing a fan on an existing installation are:

• Fan diameter.
• Installed motor horsepower.
• Gear reduction ratio of gear reducer.
• Shaft size or gear reducer model.
• Some estimate of elevation above sea level of installation.

2) Fan Selection

In selecting an optimum diameter of fan, number of fan blades, type of blades, required pitch angle, fan speed, and some estimate of horsepower via the computer aid fan model selection program the following factors should be considered and provided.

(1)Fan Use : Induced or Forced Draft Cooling Tower
(2)Airflow Volume (acfm): A typical airflow volume by the diameter of fan is as below. The normal air velocity at fan inlet is 1,600 to 2,000 fpm and the air velocity must not exceed 2,100 fpm. This is a guide line for the optimum performance of fan. Excessive airflow may result in a waste of horsepower and in a high drift loss of water droplets through the fan cylinder.

(3) Diameter of Fan (ft) or Cross Sectional Area of Tower (ft²): The fan diameter has significant bearing on performance primarily because the diameter effects the magnitude of the velocity pressure which is a parasitic loss and the fan diameter effects the pressure capability of the fan. In our estimation, velocity pressure should fall in the range of 0.14 to 0.25 inch water (1,600 to 2,000 fpm at 0.068 lb/ft³) for optimum performance. Of course, other factors influence the choice of fan diameter, such as cell-size limitations, or selection of a fan for an existing installation.

Coping with a high velocity pressure requirement is just to add blades. However, this presents a good case for converting velocity pressure into useful static pressure capability by adding velocity recovery stack (called fan stack). The solidity ratio expressed by the ratio of the total width of the fan blades to the fan circumference is a way to compare a fan's pressure capability. The higher the ratio, usually the more work the fan can do. Still another aspect of optimum fan diameter is cost. Nonstandard sizes mean special handling by the fan maker at additional cost. (Standard fan diameter: 7, 8, 9, 10, 11, 12, 13, 14, 16, 18, 20, 22, 24, 26, 28, 30, 32 in feet)

For a proper air distribution the area of fan coverage shall not be less than 30% of the cross sectional area of tower for the induced draft type of cooling tower and 40% for the forced draft type of cooling tower. Less fan diameter than these percentages will require more plenum chamber for enough mixing of air induced from the eliminators. Otherwise, the poor mixed air will generate the air turbulence, will make more vibration & noise, and will reduce the fan performance. Of course, the air will not be smoothly and evenly introduced into the fan from the drift eliminator.

(4) Actual Static Pressure (inch water): The total static pressure consists of pressure drops at the air inlet & louver, the fill, the eliminator, and the fan inlet. The pressure gain using the velocity recovery fan stack can be deducted from the total static pressure if you are using the different type of velocity recovery defined in Chapter 4 (Fan Stack, 7?of incline, 70% of efficiency, 0.08% of tip clearance, inlet shape of elliptic).

The calculation of pressure drops at the air inlet & louver, eliminator and fan inlet are so easy, while the pressure drop at fill can be read from the pressure drop curve (It depends on the air velocity, water loading, depth of fill and density of air, etc. with a specific model of fill.) The pressure drops at the air inlet & louver, eliminator and fan inlet can be generally obtained from a formula of K x 1/2 x (density x 0.1922) x (V²/115,820), where K is a pressure drop coefficient, 1g = 9.8016 m/sec² = 115,820 ft/min², and 1 lb/ft² = 0.1922 inch H₂O, whose K is in general 1.5 to 3.5 for the air inlet & louver, 2.0 to 3.0 for eliminator, and 0.1 to 0.35 for the fan inlet.

(5) Temperature at Fan (^oF) & Elevation of Site (ft) or Air Density (lb/ft³): This factor is very important in selecting the fans because of their effect on fan pressure capability and horsepower requirement through change in air density. This factor is combined into a density ratio. Hudson fan curve has a monograph for the density of dry air at 70? and sea level. The exit air of wet cooling tower is not simply a saturated air, but saturated air plus water vapor.

(6) Type of Fan Inlet Shape: Elliptical Inlet Bell (H/L = 0.15 x dia. of fan / 0.1 x dia. of fan) or what? Elliptical inlet shape proves to be ideal because the air flows along the wall with a uniform velocity and to the fan with the slightest possible turbulence. Other inlet shapes may be considered to add the extra resistance to the required duty point of the fan. The extra resistance may be determined with the aid of pressure loss factor. Also, the inlet shape influences the sound production of the fans.

(7) Venturi Height (ft) from Fan Plane to Top of Fan Stack : In the case of Wet Cooling Towers, a relatively common means of improving inlet conditions, conserving horsepower, and reducing the air recirculation into the air inlet is known as a fan stack. These stacks incorporate a slightly tapered exit cone and a well rounded inlet bell. In theory, there is a significantly reduced velocity pressure at the exit compared to the plane of the fan. Since the quantity of air is the same at both planes, the recovery of velocity pressure is converted into "static regain" which lowers the total pressure requirements of the fan, thus saving horse power. Below is a general shape of fan stack and is applicable for the actual cooling tower.

(8) Obstacle Area and Distance from Fan: (ft²/ft): The obstacles are mostly constituted by supporting beams, a gear reducer, and drive shaft, etc. These interference in the air flow due to obstacles located in the front of fan. Also, these obstacles cause a partial obstruction of the cross sectional area. As a result, extra losses occur. The losses depend on the ratio of total cross sectional area of fan casing to the obstacle area, and the ratio of the fan casing diameter to the distance from obstacles to the fan blade tip. Apart from a resistance increase in the air flow, obstacles cause an increase in sound production.

(9) Fan Speed (rpm), or motor full load speed & gear reduction ratio

(10) Fan Driver: Electric motor, speed control (single speed, two speed, or variable speed)

(11) Blade Pitch Control: Adjustable or variable pitch

(12) Number of Support Beam

(13) Minimum RF Margin (%): Details are described in Chapter 5.

(14) Required Noise Level (dB(A))

(15) Materials Construction of Tower Structure and Fan Stack: This is to be considered a maximum limit on the gear reducer itself not a normal level. Common practice in US is to use the minimum number of blades to reduce the vibration in the wood structure tower and FRP fan stack. This is to avoid dangerous air load induced pulsations on the fan cylinder. The fewer blades are occurring the higher air loads and are enlarging the more intense blade passing pulsation.

As a general rule, do not select the fans close to the limit of their BHP/blade. As bhp per blade = Brake Horse Power / Number of Blade, high blade air loading result in fatigue, vibration and noise problems. So, Hudson recommends the maximum per blade to be able to allow for fans as follows and 4 hp or less per blade is trouble free in most applications. These are guide line only and the minimum number of blades per fan would also be influenced by fan ring or fan stack rigidity or unusual loading conditions.
If the bhp/blade is excessive, you may:

• Increase the number of blades specified.
• Increase the fan diameter.
• Change to a wider chord blade type, if possible.
• Decrease the static pressure, if possible.

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