Methods of varying fan capacity in a duct system
There are four basic ways in which the capacity of a fan-duct system can be altered:
(1) Changing the fan speed
(2) Partly closing a damper in the duct system
(3) Partly closing variable position guide vanes at the inlet eye of a centrifugal fan or changing the blade pitch angle of an axial flow fan
(4) Varying the effective impeller width of a centrifugal fan.
1. The first method can be used to increase or decrease the air quantity handled,
Provided the driving motor is of adequate size. If the fan speed is increased from nx to n2 rev s“1, the pressure-volume curve for the fan shifts to a new position on the same system
Curve, the point of rating on the fan curve remaining unchanged, as indicated by the points
P] and P2 in Figure 15.19. It can be seen that the new point of intersection is P2, and the fan delivers a large air quantity, Q2, at a higher fan total pressure, plF2, through the same duct system. Conversely, a reduction in speed from n to n3 gives a fan-system intersection at point P3, the fan curve having been lowered.
For most constant volume systems the fan speed is only altered during commissioning, in order to achieve the design duty, and this is conveniently done by changing the pulleys. However, variable air volume systems must have the fan speed modulated almost continuously as the required duty varies. The most efficient way of doing this is by a frequency converter. An inverter rectifies the alternating current to a direct current supply, and then uses this to produce a smooth, variable frequency, alternating current output. Controlling the frequency of the output provides the desired speed modulation for the induction motor driving the fan.
2. If a damper in the duct system is closed it changes the position of the system characteristic. The damper is part of the system and, in common with all other system components, there is a certain pressure drop across it for the flow of a certain quantity of air past it. If, by some unspecified means, the flow of air through the damper is kept fixed,
Regardless of the position of the damper, an experimental relationship can be found between the pressure drop across the damper and its position, for a fixed airflow. Section 13.13 discusses this.
The result of partly closing a damper which forms part of a system is shown in Figure 15.26(a). It can be seen that the system curve rotates upwards in an anti-clockwise direction about the zero volume point, as the damper is closed.
Figure 15.26(b) indicates what happens when a damper is used to reduce the capacity of a fan-duct system. If the fan speed remains constant at n rev s-1 and the damper in the duct is fully open, the duty is Q at pm. In order to obtain a reduced airflow Q2, the damper is partly closed and the point of intersection moves from P to P2, which is now a different point of rating on the fan characteristic and so the fan has a different efficiency. The required capacity is obtained at a fan total pressure of plF2. The failing of the method is that it is wasteful of air power. It is however, convenient.
If the output of the fan-duct system used in example 15.9 is reduced to 2 m3 s-1 by partly closing a main damper, calculate the air power wasted across the damper.
When the fan runs at 19.1 rev s’1 the duty of the system is 2.77 m3 s-1 at 463 Pa. Reference to Figure 15.19 shows that if the air quantity is to be 2 m3 s-1 the system curve must cut the fan characteristic at 520 Pa. The system loss at 2 m3 s-1, on the other hand, is only 241 Pa (see example 15.8). The damper must therefore be closed enough to absorb the difference of 520 — 241 = 279 Pa if 2 m3 s-1 is to be delivered. The system curve rotates anticlockwise about the origin as the main damper is closed until it intersects the fan characteristic at the point P4 in Figure 15.19 at a duty of 2 m3 s_1 and 520 Pa fan total pressure, hence
Wasted air power = 2 x 0.279 = 0.56 kW
If the fan efficiency were known for a speed of 19.1 rev s-1 and a flow rate of 2 m3 s“1 the wasted fan power could be calculated by equation (15.20). In a similar way, as a filter gets dirty the pressure drop across it increases, the system curve rotates counter-clockwise, the airflow rate reduces and the fan power absorbed across the filter rises.
A system of plant and ductwork has a total pressure loss of 531 Pa, of which 90 Pa is across a clean filter, when handling 3 m3 s-1. A fan having the following pressure-volume characteristic at 8 rev s’1 is connected to the system:
MV 0 0.5 1.0 2.0 2.5 3.0 3.5
Pa 360 385 410 433 424 400 343
(a) Determine the duty achieved with a clean filter.
(b) Determine the speed at which the fan must run in order to obtain a duty of 3 m3 s-1 with a clean filter.
(c) Calculate the duty at the revised fan speed when the filter is dirty and its resistance is 180 Pa.
(a) The system characteristic with a clean filter may be established by assuming a square law:
M3 s~‘ 0.5 1.0 1.5 2.0 2.5 3.0 3.5
Pa 15 59 133 236 369 531 723
This, together with the fan data for running at 8 rev s“1 is plotted in Figure 15.27. The two curves intersect at P, yielding a duty of 2.66 m3 s“1 at a fan total pressure of 418 Pa when the filter is clean.
(.b) Using the first fan law in group A the new speed, n, required to obtain 3 m 3 s’1 is: n = 8.0 x 3.0/2.66 = 9.02 rev s_1
The intersection of the fan curve at this speed with the characteristic of the system with a
Airflow in m3 s 1
Fig. 15.27 Fan and system characteristic curves for example 15.14.
Clean filter is shown by the point P! in Figure 15.27, with a fan total pressure of 531 Pa, of which the clean filter is absorbing 90 Pa.
(c) The loss through the system without a filter is thus 531 — 90 = 441 Pa for an airflow of 3 m3 s“1, shown by the point P3 in Figure 15.27. It is seen that this system curve (without a filter) cuts the fan characteristic for 9.02 rev s~’ at P0, giving a duty of 3.23 m3 s-1 at 510 Pa. When the clean filter is installed the duty is given by the point Pj but as the system runs and the filter gets dirty the effect is to rotate the system curve in an anti-clockwise direction about the origin until it intersects the fan curve at the point P2. Ths point is vertically above the point P4 and the distance between them is 550 — 370 = 180 Pa, as Figure 15.27 shows. The duty with a dirty filter is thus 2.75 m3 s”1 at 550 Pa.
3. Variable inlet guide vanes provide a way of reducing the capacity of a fan-duct system, in a manner analogous to that discussed in section 12.13 for centrifugal compressors in refrigeration systems. Such vanes are positioned in the inlet eye of a centrifugal fan in a way that permits them to assist or retard the airflow through the impeller. Each vane is radially mounted and is hinged along a radial centre-line, being thus able to vary its inclination to the direction of airflow. The inclination may be such that the swirl imparted to the airstream changes its angle of entry to the impeller blades. This is accomplished without appreciable loss of power over a large range of the fan capacity. The result of partly closing the vanes is not the same as a throttling action. Instead, it rotates the pressure-volume characteristic curve for the fan about the origin, in a clockwise direction. Figure 15.28 shows that a duty Q2 at ptF2 can be obtained if the vanes are closed enough to shift the point of operation from P, to P2. If the vanes were capable of complete closure, then, when they were so closed, the pressure-volume curve for the fan would be the ordinate through the origin. This is a practical impossibility since leakage always occurs past the vanes.
Most axial flow fans are direct-coupled to the driving motor, which offers the choice of several fixed speeds of rotation, related to the number of pairs of poles of the motor. The required duty is then obtained by selecting an appropriate blade angle. Fans driven by a given motor shift the position of their pressure-volume characteristic curve as their blade angle is changed, somewhat like the rotation of the characteristic curve of a centrifugal fan as its inlet guide vanes close (see Figure 15.28). The efficiency of axial flow fans is much better than that of centrifugal fans with forward-curved impellers and this advantage is retained over a wide range of volumetric duties.
4. Varying the effective impeller width of a centrigugal fan. The principle of running fans in parallel (see section 15.24) can be used to vary the capacity of a forward-curved centrifugal fan. If a disc is mounted on the fan shaft, within the impeller wheel of a forward-curved fan, it divides the impeller width into two parts: a section adjacent to the inlet eye that can handle airflow and a section next to the backplate of the casing that is shielded from the inlet and can handle virtually no air. See Figure 15.30(b). It follows that if the position of the disc is adjustable the effective width of the impeller can be varied with a corresponding change in the airflow rate. The pressure-volume characteristic curve will then move as the position of the disc is altered. For example, with the disc in the midposition on the fan shaft, the fan curve could be regarded as that shown for a single fan in Figure 15.30(a), cutting the system curve at the point Pj. If the disc was then moved to make available the full width of the impeller, the fan curve would also move and cut the system curve at the point P2, in Figure 15.30(a). It follows that movement of the disc, in a direction parallel to the fan shaft, will change the effective impeller width and give a modulating control over fan capacity. For the full impeller width the fan and system curves intersect at the point P2. As the disc is moved towards the fan inlet eye, less of the impeller width is used and, at the point Pb only the front half is effective. Figure 15.30(a) shows a broken line representing in general the movable fan curve. The system curve is cut at a point Q which can be moved from P2, through Pb towards the origin at O, as the position of the disc is altered. The method is termed disc throttling.
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