**7. Fan and drive system design**

The fan must, by reason of its location downstream of the test section, pose certain design problems as outlined before. These relate to noise level and sensitivity to flow maldistribution caused by high lift or bluff models in the test section.

In general, the design methods of Wallis have been employed [111-113], together with additional experimental data [114-115]. A design utilising 100% pre-rotation has been devel‐ oped in conjunction with N.P.L. type flow straighteners so as to ensure good efficiency over a wide range of flows together with reduced possibility of stall of the cascade corner vanes immediately downstream of the fan nacelle fairing.

The location of the fan in a relatively high speed portion of the tunnel is associated with a mean rotor blade flow co-efficient of 0.56, which approaches the optimum range of flow coefficients for high fan rotor efficiency with the amount of pre-rotation employed. However, there are conflicting fan duty requirements due to the need for relatively high pressure rise and low fan noise level.

Asmaybecalculatedfromtheestimatedtunnelpressurelosscharacteristic,thefandutyrequired is 3.8 in w.g. pressure rise at a flow of 1f 122,000 CFM. The tunnel coefficient utilisation is:

(test section energy)/ (Σ circuit losses) = 1.6 to 2.3

depending on the number of screens used.

more. It appears that vanes for more than 900

40 Wind Tunnel Designs and Their Diverse Engineering Applications

speed are approximately 5 x 105

**7. Fan and drive system design**

angle iron.

Turning vane t/c ratios vary between 0.7 to 1.5%.

of each turning vane and the corresponding corner fillet.

caused by high lift or bluff models in the test section.

immediately downstream of the fan nacelle fairing.

870 and that they should be set initially at a positive angle of about 30

angle of zero relative to the tunnel centreline at exit. The selection of the value of blade spacing depends upon the application envisioned. Low turbulence tunnels require that small blade spacing be used, for example, a spacing dimension of 2 inch or 3 inch would be unreasonable. Tunnels not requiring a low 'open tunnel' turbulence level might employ spacing dimensions of 12 to 24 inches. Additional compromises to be effected are those of cost and structural integrity. Small vane spacings imply a large number of thin vanes of small chord with a resulting high cost and the possibility of vibration occurring due to relatively low vane natural frequency. Tunnels designed for low corner losses might be designed with a relatively large vane spacing and chord in order to ensure Reynolds numbers in excess of about 4 x 105

The university of New South Wales tunnel employs s/c ratios of 0.25 and 0.27 for the first and the second cascade corners increasing to 0.31 for the third and fourth corners. Blade spacings vary from 2 to 5 inches and the number of turning vanes from 41 to 33 for the first and fourth cascade corners respectively. The maximum and minimum vane Reynolds numbers at design

Because the University of New South Wales wind tunnel cross section is octagonal at all cascade corners and the vane chord is an appreciable dimension, special care had to be taken in the design of the junction between the turning vanes and the octagonal fillet so as to prevent the airstream expanding and subsequently contracting in its passage around the junction zone. The problem was solved by the manufacture of special concave and convex cross sections which were fitted in the cascade corner fillets. The shape of these special corner sections was generated so as to provide a straight line intersection normal to the vane span at the junction

All turning vanes were produced from 10 gauge (1/8 inch) mild steel plate by brake pressing. The turning vanes are set in mild steel plate supporting frames which are reinforced with

The fan must, by reason of its location downstream of the test section, pose certain design problems as outlined before. These relate to noise level and sensitivity to flow maldistribution

In general, the design methods of Wallis have been employed [111-113], together with additional experimental data [114-115]. A design utilising 100% pre-rotation has been devel‐ oped in conjunction with N.P.L. type flow straighteners so as to ensure good efficiency over a wide range of flows together with reduced possibility of stall of the cascade corner vanes

suggests that a minimum of 20 turning vanes should be used in low loss corners.

and 2.4 x 105

corners should have a camber angle of 860 to

for the first and fourth corners respectively.

to 40

with trailing edge

. Salter

These requirements have led to the selection of an 8-bladed fan rotor of 5 ft diameter, limited to a maximum tip speed of 315 ft/sec. The rotor blade chords vary from 9.9 inches at the root to 6.4 inches at the tip.

The noise spectrum from an axial flow fan can be described as consisting of two components: 'broad band' noise and 'discrete frequency' noise.

Broad band noise is attributed to two basic mechanisms: vortex shedding from blade boundary layers and interactions between the blading and random turbulence in the intake flow. The theoretical analysis of Refs 116 and 117 show that, for rotor blades operating at their design point, the vortex shedding component of broad band noise is proportional to blade relative velocity to the power 5.6 and that the intake turbulence interaction component is proportional to relative velocity to the power of 4. Reduction in broad band noise can thus be realized mainly by keeping flow velocities adjacent to solid boundaries and, specifically, blade tip velocities, to minimum values consistent with satisfactory aerodynamic performance.

Discrete frequency noise is caused by periodic aerodynamic interaction between fixed and moving blade rows. Like broad band noise, discrete frequency noise has two basic mecha‐ nisms. These are the force fluctuations on individual blades which arise from variations in mean velocity of the incoming flow. The data in Refs 116-118 indicate that interaction noise is strongly dependent upon pre-rotator-rotor axial spacing and the shape and size of the individual pre-rotator vanes. The axial spacing affects mainly the potential pressure field interaction mechanism and the vane shape, the mean velocity variation mechanism. As an example, the discussion to Ref 118 indicates that the pressure variation due to the wake persistence is still about 10% of the maximum theoretically possible at a distance equal to one stator chord downstream, for typical accelerating cascades. Experimental data seems to indicate that, consistent with satisfactory aerodynamics, interaction noise is considerably reduced by using separations between stator an rotor of three-quarters to one vane chord in conjunction with small vane areas and slender profiles.

Blade sections chosen for the pre-rotators and rotor are C4 compressor sections on circular arc camber lines [112,114-115 and 119]. These sections give high isolated aerofoil lift coefficients at angles of incidence of 30 to 40 and have a high stalling incidence. The straightener design is based on the use of the symmetrical NACA 0012 section which starts to stall at about ± 140 in the isolated aerofoil condition. Pre-rotator blades of cambered plates were considered [117] because of their comparatively low cost but were abandoned in view of their relatively poor performance under off-design conditions when compared with C4 sections.

Another parameter requiring careful selection was the choice of boss ratio as this affects the overall efficiency of the fan and tail fairing diffuser assembly. Due to the proximity of the first cascade corner, this ratio was fixed at a value of 0.4 which is less than optimum for the rotor alone.

The fan rotor blades have been stressed for centrifugal loading, torsional loads and loads due to non-coincident profile centroids and estimates have been made of the blade natural frequencies [120-121]. The fan rotor was dynamically balanced to an effective centre of gravity displacement of 3 to 5 microns.

The fan design requires a power output of 90 HP at 1200 RPM and a variety of fan drive schemes were considered. Thus a 90 HP compound wound DC motor and ancillaries that included switchgear and speed variation equipment were purchased. The Ward Leonard type speed control system proposed presented considerable difficulty in providing tunnel automatic dynamic head control. In addition, aerodynamic problems were encountered in designing the drive arrangement. A conventional shaft drive through the first cascade was at first envisaged but abandoned when it was realised that the required fairing through the cascade turning vanes caused severe blockage of a component which was already heavily loaded aerodynam‐ ically. A direct mechanical drive through a right angle bevel gearbox was next considered. However, a large fairing was needed for the drive shaft and problems were encountered in a gearbox design due to high power transmission requirements in a confined space. Alternative drive systems such as eddy-current variable speed couplings and Thyristor controlled DC drives were also investigated. All these units were costly and suffered from the same basic disadvantage that the prime movers, being large, had to be located outside the tunnel and required some sort of drive shaft arrangement through the tunnel structure to the fan rotor.

Thus the feasibility of using a hydraulic drive system was studied. This system comprises an axial piston hydraulic pump driving similar motor unit and is of the same order of cost as the other systems. The system has many advantages, the main ones being:


The system finally selected consists of a 150 HP induction motor of 92% efficiency, driving a 'Lucas' PM 3000 series, seven axial piston hydraulic pump fitted with servo-control of the swashplate angle. The servo is operated by a standard 3-15 psi diaphragm actuator. The pump provides high pressure oil at approximately 2300 psi which is supplied to, and returned from the motor by 1 ½ inch outside diameter high pressure tubes through the fan straightener and supporting vanes. Oil flow is approximately 3500 GHP and the overall efficiency of the combined pump and rotor unit is of the order of 82 to 85 %, over the complete speed range. The system also includes ancillary equipment such as a 70 gallon oil reservoir, an oil cooler, boost pump and oil filtration equipment. The main disadvantage of the arrangement is high noise level from the rotor. Provision was, therefore, included in the design for reducing noise transmission of both hydraulic pump and motor.

because of their comparatively low cost but were abandoned in view of their relatively poor

Another parameter requiring careful selection was the choice of boss ratio as this affects the overall efficiency of the fan and tail fairing diffuser assembly. Due to the proximity of the first cascade corner, this ratio was fixed at a value of 0.4 which is less than optimum for the rotor

The fan rotor blades have been stressed for centrifugal loading, torsional loads and loads due to non-coincident profile centroids and estimates have been made of the blade natural frequencies [120-121]. The fan rotor was dynamically balanced to an effective centre of gravity

The fan design requires a power output of 90 HP at 1200 RPM and a variety of fan drive schemes were considered. Thus a 90 HP compound wound DC motor and ancillaries that included switchgear and speed variation equipment were purchased. The Ward Leonard type speed control system proposed presented considerable difficulty in providing tunnel automatic dynamic head control. In addition, aerodynamic problems were encountered in designing the drive arrangement. A conventional shaft drive through the first cascade was at first envisaged but abandoned when it was realised that the required fairing through the cascade turning vanes caused severe blockage of a component which was already heavily loaded aerodynam‐ ically. A direct mechanical drive through a right angle bevel gearbox was next considered. However, a large fairing was needed for the drive shaft and problems were encountered in a gearbox design due to high power transmission requirements in a confined space. Alternative drive systems such as eddy-current variable speed couplings and Thyristor controlled DC drives were also investigated. All these units were costly and suffered from the same basic disadvantage that the prime movers, being large, had to be located outside the tunnel and required some sort of drive shaft arrangement through the tunnel structure to the fan rotor. Thus the feasibility of using a hydraulic drive system was studied. This system comprises an axial piston hydraulic pump driving similar motor unit and is of the same order of cost as the

**1.** The drive motor is only 10 inches in diameter and 20 inches long for maximum power output of 125 HP at 1200 RPM. It fits radially inside the fan nacelle fairing where the local diameter is 23 to 24 inches. This eliminates aerodynamic problems associated with a drive

**2.** Automatic tunnel dynamic head control can be obtained with conventional pneumatic

**3.** The motor speed is fully variable from 0 to 14oo RPM by means of a diaphragm actuator

**4.** The hydraulic pump can be driven by a standard 415 volt, 3-phase induction motor, for

The system finally selected consists of a 150 HP induction motor of 92% efficiency, driving a 'Lucas' PM 3000 series, seven axial piston hydraulic pump fitted with servo-control of the

performance under off-design conditions when compared with C4 sections.

other systems. The system has many advantages, the main ones being:

shaft through the tunnel structure.

and conventional pressure regulator.

control equipment to a repeatability of ± 0.4 %.

which installed electrical capacity was available.

alone.

displacement of 3 to 5 microns.

42 Wind Tunnel Designs and Their Diverse Engineering Applications

The fan and drive system and first cascade corner are mechanically isolated from the rest of the tunnel structure, and the laboratory floor, so as to prevent the possibility of any vibrations being transmitted to the test section or instrumentation.

The fan rotor is mounted on an overhung bearing assembly supported off the front of the straightener vane assembly. The straightener vanes are manufactured from ¼ inch mild steel plate with radial and longitudinal plate stiffeners which both provide torsional rigidity and define the aerodynamic profile of the straighteners. The front and the rear of the straightener vanes are attached to heavy steel diaphragm plates at the hub. The front diaphragm plate supports a rigid bearing assembly which carries the overhung fan rotor. The rear diaphragm plate carries another diaphragm plate to which is bolted the hydraulic motor. A flexible coupling connects the very short fan rotor drive shaft and the motor output shaft between the front and rear diaphragm plates. Provision is made in the rotor bearing design to absorb the 400 lb rotor thrust loading. The five straightener vanes have bolted-on cast aluminium nose and tail pieces with the sides sheathed in 16 gauge aluminium sheet.

The fan rotor is of built up construction with blades being held in split root fixings which are in turn clamped between mild steel shroud plates. The rotor blades are high quality aluminium alloy castings with large cylindrical root attachments which enable the blades to be adjusted to any angle by releasing the rotor shroud plate clamping bolts.

The pre-rotator vanes are aluminium alloy castings and are clamped between the shroud plates at the roots to form a rigid prerotator drum assembly. The nacelle nose and tail fairings are spun from 16 gauge aluminium alloy sheet. The nose fairing is bolted on to the front of the pre-rotator drum and the tail fairing to the rear diaphragm plate carrying the hydraulic motor.

Estimations have been made of the tunnel air temperature rise due to power dissipation around the circuit. It was found that without any form of tunnel air exchange or heat exchanger, the air temperature rose by as much as 10 to 150 C above ambient after a period of operation of about 10 minutes at a speed of 150 ft/s in the principal test section. This may be doubled for long periods of operation at 200 ft/sec.

The tunnel control system is reasonably straight forward. Instrumentation comprises an optical tachometer, electric drive motor anemometer and pressure gauges for hydraulic system. Electrical interlocks are provided against loss of hydraulic boost pressure and inadvertent starting of the hydraulic system with the hydraulic motor set at the maximum speed condition. Possible fan blade failures are provided for by a fan vibration cut-out switch.
