NBSIR 85-3168

Performance Requirements and

Preliminary Design of A Boundary

Layer Wind Tunnel Facility

R. D. Marshall

U S. DEPARTMENT OF COMMERCE

National Bureau of Standards

National Engineering Laboratory

Center for Building Technology

Gaithersburg, MD 20899

May 1985

4)0

100

U S. DEPARTMENT OF COMMERCE

ilATIONAL BUREAU OF STANDARDS

NBSIR 85-3168

PERFORMANCE REQUIREMENTS AND

PRELIMINARY DESIGN OF A BOUNDARY

LAYER WIND TUNNEL FACILITY

R. D. Marshall

U.S. DEPARTMENT OF COMMERCE

National Bureau of Standards

National Engineering Laboratory

Center for Building Technology

Gaithersburg, MD 20899

May 1985

U.S. DEPARTMENT OF COMMERCE, Malcolm Baldrige, Secretary

NATIONAL BUREAU OF STANDARDS. Ernest Ambler. Director

TABLE OF CONTENTS

Page

ABSTRACT vii

DESIGN OVERVIEW viii

1. INTRODUCTION 1

2. PERFORMANCE REQUIREMENTS AND DESIGN CONSTRAINTS 1

2.1 GENERAL 1 2.2 SPEED RANGE I 2.3 FLOW QUALITY 3

2.4 SIZE OF TEST SECTION 3

2.5 OTHER CONSIDERATIONS 5

2.6 SPACE LIMITATIONS 6

2.7 SUMMARY OF DESIGN CRITERIA 6

3. CLOSED-CIRCUIT TUNNEL 6

3.1 GENERAL 6

3.2 TEST SECTION 7 3.3 TURNING VANES 8

3.4 FIRST WIDE-ANGLE DIFFUSER 9

3.5 SECOND WIDE-ANGLE DIFFUSER 13

3.6 SETTLING CHAMBER 14

3.6.1 Damping Screens 15

3.6.2 Honeycomb 15

3.6.3 Arrangement of Flow Conditioners 16

3.7 EFFECT OF THE CONTRACTION 17

3.7.1 Nonuniformity of Mean Flow 17

3.7.2 Reduction of Turbulence 18

3.8 DESIGN OF THE CONTRACTION 19

3.9 ESTIMATION OF ENERGY RATIO 22

3.10 PRESSURE PROFILES 23

3.11 FAN SELECTION 24

3.12 POWER REQUIREMENTS 25

3.13 MOTOR AND CONTROLLER 2 5 3.14 SUMMARY 26

4. OPEN-CIRCUIT TUNNEL DESIGN 26

4.1 GENERAL 26

4.2 WIDE-ANGLE DIFFUSER 26

4.3 TEST SECTION 2 8

4.4 ESTIMATION OF ENERGY RATIO 28

iii
TABLE OF CONTENTS (Continued)

Page

4.5 PRESSURE PROFILES 28

4.6 FAN SELECTION 28

4.7 POWER REQUIREMENTS 29

4.8 MOTOR AND CONTROLLER 30

4.9 SUMMARY 30

5. CONCLUSIONS 30

6. REFERENCES 32

APPENDIX I - NOTATION 35

iv
LIST OF TABLES

Page

1. Contraction Geometry and Pressure Gradient 37

2. Dimensions of Three Similar Contractions 38

3. Calculation of Losses and Energy Ratio. Closed-Circuit Tunnel 39

4. Calculation of Losses and Energy Ratio. Open-Circuit Tunnel 40

v
LIST OF FIGURES

Page

1. Closed-Circuit Tunnel. Plan and Sectional Elevation 41

2. Turning Vane Profile. Closed-Circuit Tunnel 42

3. Typical Vane Cascade. Closed-Circuit Tunnel 43

4. First Wide-Angle Diffuser. Closed-Circuit Tunnel 44

5. Second Wide-Angle Diffuser. Closed-Circuit Tunnel 45

6. Settling Chamber 46

7. Contraction Profile. Definition Sketch 47

8. Pressure Coefficient and Gradient for Three Similar Contractions .... 48

9. Pressure Profiles. Closed-Circuit Tunnel. (UQ = 80 fps) 49

10. Tunnel and Fan Characteristics. Closed-Circuit Tunnel 50

11. Power Demand Curve. Closed-Circuit Tunnel 51

12. Open-Circuit Tunnel. Plan and Elevation 52

13. Wide-Angle Diffuser. Open-Circuit Tunnel 53

14. Pressure Profiles. Open-Circuit Tunnel. (UQ

= 80 fps) 54

15. Tunnel and Fan Characteristics. Open-Circuit Tunnel 55

16. Power Demand Curve. Open-Circuit Tunnel 56
ABSTRACT

This report describes performance characteristics and design details of a boundary layer wind tunnel for supporting research activities within the

Center for Building Technology. Two preliminary designs, the first consisting

of a conventional closed-circuit scheme in an over/under configuration and the

second consisting of an open-circuit scheme with a "pusher" or "blow-down"

configuration, are addressed. Both tunnels incorporate a three-dimensional

contraction with a 3.2:1 contraction ratio and a test section having a width

of 10 ft (3.048 m) and a height of 6 ft (1.829 m). The test section length

allows for the natural development of rough^wall boundary layers of sufficient

depth for 1:500 scale simulations of the atmospheric boundary layer. To ensure

stable operation over the design speed range of 1 to 75 fps (0.30 to 22.86 m/s),

both variable pitch and variable speed fan control will be required.

Key Words: Aerodynamics; boundary layers; buildings; modeling; wind

engineering; wind tunnels.

vii
DESIGN OVERVIEW

This report addresses performance requirements and the preliminary design of a boundary layer wind tunnel facility for supporting research activities that

involve the physical modeling of the planetary boundary layer and its effects on engineering structures. It is anticipated that the wind tunnel will be

supported and operated under the concept of a regional research facility.

Thus, in addition to supporting selected research activities within the Center

for Building Technology, the facility will be utilized by universities, research

establishments and other Federal agencies having a common interest in wind

engineering.

Potential research applications include, but are not limited to, the following

topics

.

o Wind forces on buildings and other structures

o Aeroelastic phenomena

o Natural and forced ventilation of buildings

o Inf iltration/exf iltration

o Thermal performance of buildings

o Modeling and measurement techniques

o Instrument development

o Interpretation and extension of full-scale measurements

o Related topics such as turbulent diffusion, smoke migration and drifting

snow.

In addition to those measurements required for the characterization of turbulent

flows, the supporting measurement capability will include pressure, force,

moment, torque, displacement, acceleration and temperature. Certain studies

may require capabilities for flow visualization, the measurement of surface

shear stress, or the measurement of tracer gas concentrations. In this design

effort the emphasis has been placed on aerodynamic issues. Wind tunnel

instrumentation and structural design issues will be addressed elsewhere.

Although a thermal modeling capability would be useful, the requirements for

such a capability within the Center for Building Technology have not been

developed to the point where they can be put in the form of wind tunnel design

criteria. Two basic tunnel designs have been considered; the first employing a closed-circuit arrangement (figure 1) and the second consisting of an opencircuit arrangement (figure 12). Both designs can meet the performance

requirements for speed range, flow quality and physical size of the test section.

However, only the closed-circuit design is adaptable to thermal modeling.

Construction costs for the open-circuit design are expected to be substantially

less than those associated with the closed-circuit design.

The laboratory space available for tunnel construction is 29 x 132 ft

(8.84 x 40.23 m) in plan and the clear headroom is 21 ft (6.40 m). Allowing

for a modest contraction ratio and adequate working space along the sides and

end of the tunnel, the test section dimensions are limited to width = 10 f

t

(3.048 m), height = 6 ft (1.829 m) and length = 60 ft (18.288 m). The range of

operating speeds is dictated by similarity requirements; the Froude number

viii
criterion indicating a minimum speed of approximately 1 fps (0.03 m/s) and the

Reynolds number criterion indicating an upper limit of from 70 to 80 fps (21.34

to 24.38 m/s) to achieve critical values of Reynolds number. The intensity of

turbulence and nonuniformity of the flow entering the test section do not place

significant constraints on the tunnel design and nominal values of 0.5 and + 1 percent have been adopted as performance criteria for intensity of turbulence

and velocity nonuniformity, respectively.

Scale ratios of from 1:50 to 1:500 constitute a practical limit on partial-depth

modeling of the planetary boundary layer on one hand and the construction of

detailed building models on the other hand. Blockage effects due to boundary

layer development and the presence of models in the test section require that

the ceiling height be adjustable in order to maintain a zero pressure gradient

along the length of the test section. Wide-angle diffusers have been employed

in both designs to allow adequate length in the test section for natural

development of a turbulent boundary layer. Use of these diffusers entails

significant head losses with a correspondingly low tunnel energy ratio. The

estimated energy ratio for the closed-circuit tunnel is 0.53, compared with 0.39

for the open-circuit design.

Both designs require a variable speed drive motor which can be of the induction

type with frequency regulation or of the direct current type with voltage

regulation. The closed-circuit design incorporates an axial flow fan with

selectable pitch blades. Pitch selection is essential for stable tunnel operation at the lower end of the design speed range. The open-circuit design

utilizes a double-wheel, double-inlet centrifugal fan with backward-curved

blades. In addition to a variable speed drive, this fan will require adjustable

inlet vanes to ensure stable operation at low speeds.

Basic dimensions of the two designs and their performance characteristics are summarized in the following table.

ix
SUMMARY OF BASIC DIMENSIONS AND PERFORMANCE CHARACTERISTICS

FOR TWO TUNNEL DESIGNS

Feature Closed-Circuit Tunnel Open-Circuit Tunnel

General arrangement Over/under Blowdown or "pusher'

Overall length

Test section dimensions

117 ft 111 ft

Width 10 ft 10 ft

Height 6 ft 6 ft

Length 55 ft 50 ft

Speed range

Speed deviation across

1 to 7 5 fps 1 to 7 5 fps

test section + 1 percent + 1 percent

Turbulence intensity 0.5 percent 0.5 percent

Flow conditioning Honeycomb/screen Honeycomb/screen

Contraction ratio 3.2:1 3.2:1

Fan type Axial flow Centrifugal; double

inlet, backwardbladed

Blade or wheel diameter 96 inches 98 inches

Power 150 hp 200 hp

Speed regulation Variable speed drive

and selectable pitch

blades

Variable speed driv<

and vaned inlet

Energy ratio 0.53 0.39

1 ft = 0.3048 m

x
1. INTRODUCTION

This report describes the development of performance criteria and the

preliminary design of a closed-circuit boundary layer wind tunnel and an

alternate open-circuit tunnel for supporting selected research activities

within the Center for Building Technology. The design effort described herein

addresses aerodynamic considerations and power requirements only. Detailed

structural design and instrumentation requirements will be addressed elsewhere.

The major function of a boundary layer wind tunnel facility will be to support

wind engineering research activities. It is anticipated that wind tunnel

investigations will range from studies of the aeroelastic behavior of bluff

bodies placed in smooth flow to assessments of the dynamic response of complex

engineering structures immersed in highly turbulent boundary layers. The

development of thick, turbulent boundary layers as representations of the

atmospheric boundary layer will also be an area of study.

In addition to measurements normally required to characterize the flow in a wind tunnel test section, measurements of the steady and fluctuating components

of wind-induced pressures, forces, moments and displacements will be carried

out. Improved instrumentation and measurement techniques are expected to be

major subjects for study. These activities will support the traditional role

of NBS/CBT in measurement technology and assurance.

Other areas of research within the Center for Building Technology that are

expected to benefit from a boundary layer wind tunnel facility include natural

and forced ventilation, inf iltration/ exfiltration, building heat transfer,

drifting snow and diffusion of pollutants. Potential applications outside the

Center for Building Technology include studies of smoke movement in buildings

by the Center for Fire Research and building research needs of other Federal

agencies for which the simulation of the atmospheric boundary layer is a requirement

.

2. PERFORMANCE REQUIREMENTS AND DESIGN CONSTRAINTS

2.1 GENERAL

The following sections of this report identify the performance characteristics

that a boundary layer wind tunnel should possess in order to meet the needs of

the anticipated research applications. Limitations of cost and available

laboratory space are also considered in arriving at a set of criteria for the

aerodynamic design of a boundary layer wind tunnel facility.

2.2 SPEED RANGE

There are a number of factors that influence the choice of wind tunnel speed

range which ultimately involves a compromise between similarity requirements

and constraints such as available space, size of facility and cost. Considering

first the investigation of wind forces acting on buildings and similar structure, ,

1
which will be the primary function of the facility, the basis for similitude

should be the Reynolds number

Re Ud

v

( 1 )

where U is the wind speed, d is a characteristic dimension (such as breadth or diameter) of the structure and v is the kinematic viscosity. Taking 1:50 to

1:500 as a practical range of geometric scale ratios for modeling buildings and

other structures, Reynolds number equality would require tunnel speeds of from

50 to 500 times the full-scale speeds. Fortunately, the Reynolds number

criterion usually is not important in the case of sharp-edged bodies for which

the points of flow separation tend to be fixed. However, this relaxation of

the Reynolds number criterion cannot, in general, be extended to bodies with

curved surfaces, e.g. , domes and circular cylinders.

With regard to studies involving isolated bluff bodies such as circular

cylinders, the facility should be capable of attaining critical Reynolds

numbers as this flow regime is encountered in many practical situations and is

basic to the understanding of fluid/structure interaction. Critical Reynolds

numbers for a circular cylinder range from about 2 x 10^ to 6 x 10^, depending

on the incident turbulence and the surface roughness [1]. By choosing a sufficiently large cylinder diameter, this range of Reynolds number could be

attained at modest tunnel speeds. However, as will be discussed later, the

effects of blockage place rather severe limitations on the size of models that can be tested in an enclosed jet and a cylinder diameter of 0.5 ft (152 mm)

can be taken as a reasonable upper bound. The corresponding range of tunnel

speeds required to attain critical Reynolds numbers is approximately 70 to

200 fps (21.34 to 60.96 m/s). Since the power demand is proportional to the

cube of the speed, there is little incentive to set the upper bound on speed

much above the lower end of this range.

Two other dimensionless numbers that can be of importance in the modeling of

certain engineering structures are the reduced velocity

Ur unQ

d

and the Froude number

Fr -

U

V'gd

( 2 )

(3)

were iIq is the frequency associated with a particular mode shape of the

structure and g is the gravitational constant. Since the frequency, tIq , can

be altered by changing the construction of the model, the reduced velocity is

not a major consideration in establishing the speed range of the tunnel.

2
However, Froude number equality requires relatively low tunnel speeds; approximately 4 percent of the prototype speed for a scale ratio of 1:500. It is

not unusual for certain full-scale structures to exhibit significant dynamic

response at speeds of 20 mph (8.94 m/ s) or less. A 1:500 scale simulation

modeled on the basis of the Froude number will require stable tunnel operation

at speeds as low as 1 fps (0.30 m/s).

Based on the factors discussed in the preceding paragraphs, a speed range of 1 to 80 fps (0.30 to 24.38 m/s) is a reasonable choice for the investigation of

wind forces on buildings and other structures. This precludes the modeling of

most engineering structures on the basis of Reynolds number equality, a problem

that is common to practically all low speed boundary layer wind tunnels.

Similitude requirements for other modeling applications such as drifting snow

and forced convection place certain demands on tunnel speed, but these demands

tend to fall within the range required for the investigation of wind forces

and structural response. Modeling laws for many non-structural applications

can be found in references 2 and 3.

2.3 FLOW QUALITY

Flow quality includes background turbulence and uniformity of the flow entering

the test section. Neither of these factors poses a serious limitation on

studies conducted in thick, turbulent boundary layers. However, background

turbulence and flow uniformity can be critical when conducting studies in

smooth flow, the aeroelasticity of bluff bodies being one example.

Performance characteristics for a number of low-speed wind tunnels specifically

designed for industrial aerodynamics applications are listed in reference 4.

These characteristics suggest that turbulence intensities of the order of 0.5

percent and nonuniformity of the flow of about 4- 1 percent on entry to the test

section can be achieved using a modest number (2 or 3) of damping screens and

contraction ratios (ratio of inlet area to outlet area) that range from 3:1 to

8:1. Since these levels of turbulence intensity and nonuniformity of flow

present no serious limitations on the research applications envisioned for the

tunnel, they will be adopted as design criteria.

2.4 SIZE OF TEST SECTION

In order to set limits on the size of the test section, it is necessary to

establish the model scale and the amount of blockage that can be tolerated.

Issues that relate to the selection of model scale in the modeling of engineering structures and atmospheric boundary layers have been assessed by Vickery

[5] and by Teunissen [6]. For large engineering structures and full-depth

simulations of the atmospheric boundary layer, Vickery concludes that the

linear scale should not be smaller than about 1:500 and that the practical

consideration of being able to work in the test section while standing up sets

the lower limit on height of the test section at about 6 ft. (1.83 m).

3
The minimum width of the test section can be related to the diameter of the

turntable required for mounting models of adjacent structures and significant

terrain features. Ideally, the radius of the turntable should be as large as the distance over which wake effects produced by objects surrounding the test

structure are likely to persist. Britter [7] has addressed this question and

suggests that the maximum velocity deficit, AUjq^x, behind a bluff body immersed

in a turbulent boundary layer decays according to the relationship

where AU is the difference between the approach flow velocity and the wake

velocity, is the approach flow velocity at the height h of the bluff body,

A is a function of the body geometry and boundary layer depth, and X is the

distance downstream of the body. Measurements carried out on models placed in

a turbulent boundary layer suggest that A - 2.5 for typical building shapes.

Assuming a prismatic building 100 x 100 ft (30.48 x 30.48 m) in plan, 400 ft

(122 m) high and modeled at a linear scale of 1:500, approximately 7 ft (2.13

m) would be required for the velocity deficit to decay to an acceptable level

of approximately 10 percent of the approach flow velocity. This implies a minimum turntable diameter of about 14 f t (4.27 m) and a slightly larger test

section width. The 14 ft (4.27 m) requirement would enable a complete model

of surrounding bluff bodies to be mounted conveniently on the turntable.

However, this is a convenience rather than a strict requirement for the

modeling of significant surrounding features of a given full-scale site.

Other factors that influence choice of test section width are the effects of

blockage, available laboratory floor space, capacity of the fan and required

power. With regard to blockage effects, a solid blockage of from 3 to 5 percent

of the test section area should be considered as an upper limit [5,8] if significant distortions in surface pressures and drag forces are to be avoided.

Taking as the limiting case a circular cylinder spanning the test section ver- tically, a test section width of 10 ft (3.048 m) would be required to satisfy the

5 percent blockage limit for a cylinder diameter of 0.5 ft (152 mm). As was noted earlier, this is the diameter required to attain critical Reynolds numbers

when operating at maximum tunnel speed. Also, a width in excess of 10 ft

(3.048 m) would seriously limit the laboratory working space available along

the sides of the test section. Finally, a tunnel speed of 80 fps (24.38 m/s)

and test section dimensions of W = 10 ft (3.048 m) and H = 6 ft (1.829 m) will

require a flow rate of 288,000 cfm (136 m^/s). Axial flow fans with capacities

up to about 300,000 cfm (142 m^/s) and of standard design are readily available

while fans of larger capacity tend to be special designs with a correspondingly

higher cost. For these reasons, 10 ft (3.048 m) is considered to be a reasonable

choice for width of the test section.

The length of the test section is dictated by the distance required to develop

a turbulent boundary layer of sufficient thickness to simulate atmospheric

boundary layers. Assuming a full-scale thickness of approximately 1500 ft

(457 m) the corresponding thickness in a 1:500 scale simulation would be 3 ft

aumax A

( 4 ) (X/h) 3/^ 2

4
(0.91 in). Thick, turbulent boundary layers can be developed in two ways;

(1) by providing sufficient length in the test section for natural development

over a suitably rough surface; or (2) by installing devices at the test section

entrance that will create an initial momentum deficit near the floor of the

tunnel and that will accelerate the growth of the layer over a rough boundary.

Both methods are of interest in developing techniques for modeling atmospheric

boundary layers and, therefore, the method of natural development will dictate

test-section length. Measurements of the development of rough-wall boundary

layers having length scales of the order being considered here suggest a development length of approximately 60 ft (18.29 m) [9]. This is the distance

from the test section entrance to the location of the model under test.

Additional space will be required between the model and the end of the test

section to allow for wake development and decay.

2.5 OTHER CONSIDERATIONS

It is conceivable that certain research activities will require the modeling of

heat transfer phenomena or the effects of thermal stratification on boundary

layer flows. Thermal modeling has not been addressed in the current design

effort for two reasons. First, the general requirements for a thermal modeling

capability within the Center for Building Technology have not been finalized.

Second, construction costs for a boundary layer wind tunnel with state-of-theart thermal modeling capabilities such as those described in reference A are

estimated to be from three to four times the costs associated with a conventional facility. Nevertheless, the closed-circuit design described in this

report is, in principle, adaptable to thermal modeling.

Consideration has also been given to the acoustic environment in the sense that

provision has been made for the installation of silencers and the head loss due

to these devices has been included in the estimates of the tunnel energy ratio.

In addition to its adverse effect on the general working environment, noise

produced by the drive section and return duct can cause very serious problems

with the measurement of surface pressures on models under test. Because the

fan requirements are only approximately established in this report, actual

design of the silencers will be carried out after a specific fan has been

selected. Finally, tunnel cooling may be an important factor, depending upon

the tunnel configuration. If the closed-circuit arrangement with a thermal

modeling capability is selected, tunnel cooling will be a part of the thermal

capability. However, if a conventional closed-circuit arrangement is adopted,

provision must be made to remove heat from the tunnel. Otherwise there will

be serious problems with instrument drift and a very uncomfortable working

environment in the test section. A feature of the closed-circuit design

developed here is that space has been provided in the low-speed region of the

tunnel for the installation of a heat exchanger. Cooling is not expected to

be a serious problem with an open-circuit tunnel since normal air exchange in

the space occupied by the tunnel will be sufficient to handle the heat gener.iii-d

under most operating conditions.

5
2.6 SPACE LIMITATIONS

The space proposed for housing the boundary layer wind tunnel facility is

located in Building 202 on the NBS Gaithersburg site. At the present time this

space is occupied by a number of small lab modules that were installed in what

was originally a large open floor area. It is proposed that this open space be

restored by removing a number of concrete masonry partition walls and a portion

of the concrete floor slab installed at the second-floor level during the

conversion to small lab modules.

The proposed renovation will provide a clear floor area of 29 x 132 ft (8.84 x 40.23 m) and a clear headroom of 21 ft (6.40 m). The roof structure consists

of a metal deck placed over bar joists with a depth of 3 ft (0.91 m). This is

of significance in the case of an open-circuit tunnel because of the additional

area provided by the bar joists for the return flow. The floor consists of a

6 in (152 mm) concrete slab on grade. Removal of part of this slab will be

required for a closed-circuit arrangement.

2.7 SUMMARY OF DESIGN CRITERIA

The design criteria developed up to this point can be summarized as follows:

Test section Dimensions:

Width = 10 ft (3.048 m).

Height = 6 ft (1.829 m).

Length = 60 ft (18.288 m) to turntable.

Speed Range:

Continuously variable from 1 to 80 fps (0.30 to 24.38 m/s).

Flow Quality:

Intensity of background turbulence =0.5 percent.

Maximum nonuniformity = + 1 percent.

Space Available for Construction:

Width = 29 ft (8.84 m)

Length = 13 2 ft (40.23 m)

Height = 21 ft (6.40 m)

3. CLOSED-CIRCUIT TUNNEL

3.1 GENERAL

The tunnel arrangement shown in figure 1 is believed to be a reasonable compromise

among performance requirements on the one hand and constraints due to available

6
space, ease of fabrication and cost of construction on the other hand. The

following sections of this report describe the design considerations that

underlie the tunnel layout and the special features shown in figure 1.

3.2 TEST SECTION

The arrangement of the test section in figure 1 meets the general requirements

for flow area and minimum working length prescribed by the performance criteria.

A 12 ft (3.66 m) clearance is provided between the end of the tunnel and the

laboratory wall line so that equipment can be moved to either side of the test

section. Allowing room for the first and fourth corners, the wide-angle

diffuser, the settling chamber and the contraction, 69 ft (21.031 m) are available for the test section. Sufficient space must be provided between the

turntable and the first corner to allow room for wake development behind the

models under test. This has been set at 14 ft (4.267 m) or approximately 20

model diameters, leaving 55 ft (16.764 m) between the test section entrance and

the center of the turntable for boundary layer development. A second turnable

is located near the test section entrance for conducting studies in smooth,

uniform flow.

Nominal height of the test section is 6 ft (1.829 m), but this can be varied

within the limits indicated by the dashed ceiling profile lines shown in figure

1. For the case of an empty tunnel the development of a turbulent boundary

layer is given by the relationship

0.376 X

<R

e

) 1/5

(5)

where 6 is the boundary layer thickness, X is the development length or distance

downstream of the boundary layer origin and Rg is the Reynolds number based on

the free-stream velocity and the development length [10]. The blockage effect

of the boundary layer is given by the displacement thickness

6* = ( 6 )

It is seen that the depth of the boundary layer and the displacement thickness

increase with decreasing tunnel air speed. For a speed of 10 fps (3.05 m/s), the

boundary layer depth and displacement thickness at the turntable will be 1.03

and 0.13 ft (314 and 40 mm), respectively. Boundary layer development occurs

on all four surfaces of the test section and, therefore, the net blockage is

approximately (0.13) (perimeter) = 4.2 ft^ (0.390 m^). The corresponding ceiling

adjustment required to maintain a zero longitudinal pressure gradient along the

length of the test section will be 4.2/10 = 0.42 ft (128 mm).

The ceiling adjustment required for a rough-wall boundary layer can be estimattd

from the power law representation of the mean velocity profile for an urban

exposure [11]

7
u 0.33

z (7)

Uoo

where Uz is the mean velocity at height z and Uoo is the gradient or free-stream

velocity. The corresponding displacement thickness is

Assuming a full-scale value of 1600 ft (488 m) for 6 and a scale ratio of

1:400, the displacement thickness is (0.248) (1600) /400 = 1.0 ft (305 mm).

Blockage due to models immersed in a turbulent boundary layer is more difficult

to assess, but the net effect will be an increase in the required ceiling

adjustment. For this reason a maximum adjustment of 2 ft (610 mm) has been

provided over the turntable. This range of adjustment also provides for a smooth transition from the 10 x 6 ft (3.048 x 1.829 m) test section to the 10 x

8 ft section (3.048 x 2.438 m) at the entrance to the first corner.

3.3 TURNING VANES

In order to limit losses to an acceptable level and maintain uniformity of flow,

it is necessary to install turning vanes or cascades in the right-angle corners

of closed-circuit tunnels. Experimental studies carried out by Salter [12] and

by Winter [13] indicate that the most efficient turning vane arrangement consists

of thin vanes fabricated from sheet metal and spaced so that the gap/ chord

ratio is approximately 0.25. As used here, "gap" refers to the vane spacing

measured along the diagonal of a right-angle corner. Another commonly used

measure of vane spacing is the pitch which is measured normal to the tunnel

axis. Thus gap = /2~~(pitch).

The vane profile recommended by Winter is shown in figure 2. The corresponding

loss coefficient was found to be 0.033 at R0 = 1.9 x 10^. This does not include

secondary flow losses at the vane roots (where the vanes intersect the tunnel

walls) so a slightly higher loss coefficient, say 0.05, would be appropriate

for design. It is expected that the performance of the turning vanes in the

first corner will be adversely affected by the presence of a thick, turbulent

boundary layer in the test section, or by the wake developed behind large bluff

bodies tested in uniform flow. Because of this a higher loss coefficient, say

0.10, should be assumed for the first corner.

Specification of a gap/ chord ratio does not, by itself, establish the chord

length or gap dimension. Measurements of losses in rectangular ducts with

short-radius bends [14] suggest that the aspect ratio (W/D where D is the radial

dimension) should not be less than 3 under any circumstances and, preferably,

should be greater than 6. Most of the closed-circuit tunnels described in

reference 4 have cascades so proportioned that the aspect ratio of the slot

between adjacent vanes ranges from 10 to 30. For the over/under closed-circuit

arrangement shown of figure 1, the section width and height at the first corner

6* = 0.248 6 ( 8 )

8
are 10 and 8ft (3.048 and 2.438 m), respectively. Fifteen vanes with a pitch

of 6 in (152 min) would correspond to a vane gap of 6 /2 = 8.49 in (215 mm).

Thus the optimum chord length would be 4 x 8.49 = 33.94 in (862 mm) and the

aspect ratio of the cascade slots would be 10 x 12/8.49 = 14.1. If the same vane gap is used in all four corners, the aspect ratio will reduce to 11.3 at

the second corner and will increase to 17.0 at the third and fourth corners.

These values are higher than is necessary to ensure good vane performance and,

because of the lower speeds, it is likely that the gap/chord ratio for the

third and fourth corners could be increased to about 0.5 without flow separation and stalling of the vanes. Selection of a smaller chord dimension would

reduce the scale of disturbances produced by the turning vanes at the fourth

corner and thus promote more rapid decay. However, the use of more than one chord and gap dimension would likely increase fabrication costs and would not

lead to a significant improvement in quality of flow in the test section since

a honeycomb and damping screens will accomplish the necessary flow conditioning.

A typical vane cascade with optimum chord length is shown in figure 3.

Turning vanes of either variable or constant thickness are commercially available

and it is doubtful that the fabrication of special vanes will prove to be

economical. However, the discussion presented here can be used as a basis for

turning vane evaluation and selection.

3.4 FIRST WIDE-ANGLE DIFFUSER

In a conventional closed-circuit tunnel, the drive section would be followed by

a short round-to-square transition and a long plane-walled diffuser of square

or rectangular cross section with a divergence angle (total included angle) of

approximately 6 degrees. The length and aspect ratio of the diffuser would be

so proportioned to make the tunnel cross section at the third corner comply

with the pre-contraction tunnel section, in this case 16 x 12 ft (4.877 x 3.658 m). This would provide maximum flow area and would minimize losses through

the third and fourth corners. Assuming an 8 ft (2.438 m) diameter for the

drive section, the mininum diffuser length that can satisfy the limitation on divergence angle is approximately 76 ft (23.17 m). When combined with the

length of the square-t o-r ound transition, the required separation between the

drive section and the entrance to the third corner is approximately 84 ft

(26.60 m).

Although there is room to accommodate a 6-degree plane-walled diffuser,

clearance at each end of the tunnel would be minimal and the cost of construction using either a lined pit or a cut-and-cover arrangement for the return

circuit would be prohibitive. An attractive alternative is to use buried

precast concrete pipe, a round-to-square transition and a wide-angle diffuser

as indicated in figure 1. This will result in reduced operating efficiency,

but the reduction in construction costs should be substantial.

The designation of "wide angle" as used here is somewhat arbitrary. Diffusers

with divergence angles substantially greater than 8 degrees are often referred

to as "wide angle" or "rapid expansion" diffusers. Although the divergence

angle is only one of several parameters that determine diffuser performance.

9
there is general agreement that divergence angles should not exceed 7 or 8

degrees unless special measures are taken to guard against flow separation and

stalling. The arrangement being considered here has a divergence angle of

22.6 degrees.

The diffuser expansion need not be limited to one plane and, in fact, it would

be possible to make the transition from an 8 x 8 ft (2.438 x 2.438 m) section

to a 16 x 12 ft (4.877 x 3.658 m) rectangular section using a wide-angle diffuser

with a two-plane expansion. The problem with this approach is that the width

and depth of the excavation could introduce additional construction problems

such as foundation underpinning. In addition, the length required upstream of

the settling chamber for the fourth corner would be approximately the same as

the arrangement shown in figure 1 for which a second wide-angle diffuser is

required to make the transition from the fourth corner to the settling chamber.

The arrangement shown in figure 4 is a plane-walled diffuser with single-plane

expansion. The area ratio (outlet/ inlet) is 1.50 and the ratio of length to

inlet width is 1.25. According to the diffuser performance data presented in

reference 15, this diffuser will exhibit transitory three-dimensional flow

separation and will operate just outside of the steady two-dimensional separated

flow regime. Clearly, this arrangement would result in low diffuser efficiency

and poor tunnel performance, the most serious effect being unsteadiness or surging in the tunnel speed.

Instabilities in wide-angle diffusers are usually associated with flow

separation at some point along the diffuser wall where the momentum in the

boundary layer is insufficient to overcome the boundary shear and the adverse

pressure gradient. Techniques that have been used to cure transitory flow

separation and stall in diffusers include removal of the boundary layer through

suction slots and the addition of momentum to the boundary layer with devices

such as vortex generators. The best solution is to avoid problems with separation and stall through proper choice of diffuser geometry.

Schubauer and Spangenberg [16] demonstrated the effectiveness of diffusing

screens in preventing boundary layer separation and in restoring separated

flows in diffusers with very large angles of divergence. Moore and Kline [17]

showed that good performance can be obtained from two-dimensional wide-angle

diffusers by using flat-plate splitter vanes installed close to the diffuser

throat. This technique was refined by Cochran and Kline [18] and a threedimensional version has been used successfully in at least one large boundary

layer wind tunnel of recent construction [4]. Whether one chooses diffusing

screens or vanes, the resulting losses in the diffuser will be greater than

would occur in a conventional diffuser without flow separation. This is the

price to be paid for improved diffuser performance.

An additional benefit of diffusing screens is the reduction of turbulence since

they also act as damping screens. Because of this, diffusing screens are usually Installed ahead of the settling chamber in blowdown tunnels (i.e.,

open-circuit tunnels with the drive section located ahead of the test section).

Their main disadvantage is that, for a given diffuser, the losses are usually

higher than those produced by vanes. Vanes are usually cheaper and easier to

install, but their greatest advantage in the application being considered here

10
is that they can also serve as acoustic silencers. An alternative would be to

install annular silencers at the fan outlet with resulting losses considerably

higher than those produced by a wide-angle diffuser fitted with vanes.

Based upon the findings presented in references 17 and 18, it is possible to

establish a three-dimensional vane arrangement (see figure 4) that can be

expected to perform adequately. However, since definitive design criteria are

not available, performance should be confirmed using a scale model of the

diffuser that includes the circular duct upstream of the diffuser and the

turning vanes in the third corner directly downstream of the diffuser. Also,

the prototype should be designed to allow for fine adjustment of the diffuser

vanes because of uncertainties with respect to scale effects.

For purposes of design it is necessary to estimate the energy loss through the

diffuser and this requires some knowledge of its performance characteristics.

Various parameters have been used to describe diffuser performance, one of the

more common parameters being the nondimensional static-pressure recovery

coefficient

C pr

p - p

_1 _ 1

_ 1/2 pUl

(9)

where the subscripts 1 and 2 refer to the diffuser inlet and outlet sections,

respectively. Neglecting for the moment such problems as nonuniform distribution of velocity and pressure, and assuming no energy losses, the theoretical

maximum value of the static-pressure recovery coefficient for a diffuser with

inlet area A]^ , and outlet area A2 is

( 10 )

Thus, for the wide-angle diffuser arrangement shown in figure 4, the upper

limit on the static-pressure recovery coefficient would be 1 - (64/96)^ = 0.56.

Equation 9 is valid only if the flow is incompressible, the flow at the diffuser

inlet is uniform and axial, and the distribution of the static pressure across

the diffuser outlet is uniform. While the flow being considered here can be

treated as incompressible, the swirl component imparted by the fan to the

approach flow and the boundary layer development along the circular duct and

circular-to-square transition will contribute to a nonuniform velocity distribution at the diffuser inlet. To account for this nonunif orraity , a kinetic

energy coefficient, a, and a potential energy coefficient, 0, are introduced

11
where

a

and

= ^ IA

(1/2 p

U

2 )UdA/ (1/2 pu|)

= —— / U

3dA

AUI A

8 “ XBT VmA/P

A

( 11 )

( 12 )

The loss in total pressure, Pt , over the length of the diffuser can be expressed

as

p tl - p

t2

- “1 d/2 P«A1>- a2

(

1 /2 PUJ)" (hrA2- Bl p

Al) 03)

where the subscript A denotes an area-averaged quantity. The loss in total

pressure is usually expressed in the form of a total-pressure loss coefficient

*t "

tl - P

t2

a-L (1/2 pUAf)

(14)

Reference 15 presents contours of static-pressure recovery coefficients and

tentative values of total-pressure loss coefficients for diffuser geometries

similar to that shown in figure 4. Suggested values for Cp r and Kt are approximately 0.40 and 0.18, respectively.

Although reference 18 does not consider diffuser length ratios (length/ inlet

width) of less than 4 and in most cases works with length ratios of 8, the data

for divergence angles of from 20 to 30 degrees indicate that a properly designed

diffuser with vanes should operate at about 80 percent of the theoretical upper

limit on Cp r. This compares with an average value of about 55 percent for a wide-angle diffuser without vanes. With regard to the total-pressure loss

coefficient, reference 18 indicates Kt = 0.15 to 0.20 for a vaned diffuser over

the same range of divergence angle. For diffusers without vanes, Kt = 0.25 to

0.55. When compared with the values of Cpr and Kt suggested in reference 15,

the addition of vanes to the diffuser shown in figure 4 would be expected to

improve Cp r slightly (0.8 x 0.56 = 0.45 vs 0.40) and Kt would be expected

to remain about the same (0.15 to 0.20 vs 0.18). Thus the major advantages of

a vaned diffuser in this case are the prevention of transitory threedimensional flow separation and the use of the vanes as acoustic silencers.

In estimating pressure profiles and power requirements, values of Cpr = 0.40

and Kt = 0.18 have been assumed.

12
3.5 SECOND WIDE-ANGLE DIFFUSER

The floor profile for diffuser No. 2 connecting the fourth corner with the

settling chamber is shown in figure 5. The approach taken here is to eliminate

the potential for flow separation by adding diffusing screens as described in

reference 16. While a vaned diffuser would be more efficient, vanes would

generate wakes directly ahead of the settling chamber and this would require

damping screens with higher combined loss coefficients to produce the same quality of flow at the entrance to the test section.

The principle underlying the selection and placement of diffusing screens in a wide-angle diffuser is to forestall flow separation by reducing the adverse

pressure gradient over the length of the diffuser. In effect, the diffuser and

screens are proportioned so as to make the static-pressure recovery coefficient

equal to zero. This can be accomplished by using a single diffusing screen

with a high loss coefficient, or by using multiple diffusing screens with lower

loss coefficients. However, the spacing of the screens along the length of

the diffuser must be close enough to prevent flow separation from occurring

along the diffuser walls between the screens.

In reference 16 the diffuser efficiency is defined as the ratio of ^ain in

potential energy to loss in kinetic energy. Thus the efficiency, E,

between any two sections 1 and 2 of the diffuser is

E

1 *2 "

3 2

PA2 3 1PA1

a

i (1/2 PU

A1

)_ a

2 (1/2 pUA2 }

( 15 )

If diffusing screens with loss coefficients and K2 are installed at sections

1 and 2, respectively, and if their effect is to make P and U uniform over each

section, the diffuser efficiency from the downstream side of screen 1 to the

downstream side of screen 2 will be

K„

A

2

2

(XI>

- 1

(16)

The wide-angle diffuser shown in figure 5 has an area ratio of 2 and horizontal

and vertical divergence angles of 36.9 degrees. There is very little information

in the literature regarding the performance of plane-walled diffusers with a two-plane expansion and it is necessary to infer certain characteristics from

diffusers with single-plane expansions. Using equations (9) and (10) and

assuming that the diffusing screens will be fully effective in promoting uniform

flow throughout the diffuser, the efficiency, exclusive of screen losses, can

be expressed as

( 17 )

13
Therefore, it will suffice to estimate Cp r

. If it is further assumed that the

static-pressure recovery coefficients suggested by reference 18 are^ typical of

wide-angle diffusers operating without separation, then Cpr - 0*8 Cp r and

E - 0.8. Setting E = 0 in equation (16), the required screen loss coefficient

for a given screen spacing is

K

x = 0.8 (18)

To reduce the likelihood of flow separation at the juncture of the fourth- corner outlet section and the wide-angle diffuser section, a smooth transition

is provided by the circular arc with 6 ft (1.829 m) radius as shown in figure

5. To simplify fabrication the transition is made an integral part of the

fourth-corner assembly and the plane of connection with the diffuser is a convenient location for the first diffusing screen. From equation (18) the

required loss coefficient is K = 0.23. If the second diffusing screen is

placed at the end of the diffuser, a loss coefficient of K = 1.70 will be

required.

Although this arrangement will satisfy the requirement that E = 0, it does not

guarantee the absence of flow separation along the diffuser walls. The most

likely location for separation to occur is along the corners of the diffuser

downstream of the first screen. If an additional screen is installed 2.5 ft

(762 mm) upstream of the diffuser outlet, the respective loss coefficients for

the first, second and third diffusing screens will be 0.23, 0.68 and 0.56. As

a part of this design effort, preliminary studies carried out on a diffuser of

similar geometry indicate that three screens are sufficient to suppress flow

separation. Screens having identical K factors may be used so long as K is not

less than 0.68. A logical choice for uniform screen type would be 20 mesh (20

wires/in) and wire diameter of 0.010 in. (0.25 mm) which is readily available

in widths of 12 ft (3.66 m) or more and for which K = 0.70.

It is of interest that Schubauer and Spangenberg [16] explored the possibility

of using curved diffuser walls of optimum profile while carrying out their

studies of diffusing screens in conical diffusers. Theoretical and experimental studies of optimum wall profiles for axisymmetric diffusers are reported

in reference 19. With optimum wall profile and careful design and location of

a single diffusing screen with nonuniform loss coefficient, it is possible to

achieve moderate levels of pressure recovery and highly uniform flow at the

outlet of axisymmetric diffusers having very small length/diameter ratios.

Although it is less efficient, the plane-walled diffuser shown in figure 5 is

relatively simple to fabricate.

3.6 SETTLING CHAMBER

The settling chamber provides space for flow modification or conditioning

devices located directly upstream of the wind tunnel contraction. Since the

contraction also functions as a flow modification device and influences the

14
flow for some distance upstream, the design of the settling chamber and the

design of the contraction are interrelated.

3.6.1 Damping Screens

The most common pre-contraction flow conditioning device is the damping screen.

Its function is to accelerate the decay of incident turbulence and promote

uniformity of the flow. On the basis of extensive tests concerned with the

modification of turbulence intensity by wire mesh screens, Dryden and Schubauer

[20] suggested that the mean intensity of turbulence in a flow passing through

a screen is reduced by the ratio

1

/I + K

(19)

where 1^ is the turbulence intensity upstream of the screen, I2 is the downstream

intensity, and K = the head-loss or pressure-drop coefficient. The value of K

is dependent upon Reynolds number, but may be assumed to be constant for screens

with solidity ratios (solid area/total area) of less than about 0.6 and with

Re > 150. For Re < 7 5, K increases rapidly with decreasing speed.

In view of the form of equation (19), obviously it is more efficient to use n screens with head-loss coefficient K than it is to use one screen with headloss coefficient nK to achieve the desired reduction in turbulence intensity.

While there is a minimum spacing of screens required to produce maximum damping

of turbulent fluctuations (400 to 500 wire diameters), screen spacing is usually

dictated by the working space required for cleaning the screens. Thus the

screen spacing should not be less than about 2 ft (610 mm).

With regard to nonuniformity of flow, it can be shown [21] that if U]_= + u^

is the local longitudinal velocity far upstream and U

2 = + U

2 is the corresponding value far downstream of the screen, where the subscript A denotes

averaging over the tunnel cross section, then

^2 =

2 - K

ux 2 + K

( 20 )

so that nonuniformity of flow is completely removed by a screen with K = 2.

More recent work and measurements carried out in actual wind tunnel installations suggest that the optimum head-loss coefficient for the reduction of spatial

variations in velocity is approximately 2.8 [5,22].

3.6.2 Honeycomb

Although wire mesh screens can be very effective in damping turbulent

fluctuations and in reducing non-uniformity of flow, they are not particularly

15
effective in removing steady transverse components of velocity induced by the

fan in closed-circuit tunnels or by the inlet in open-circuit tunnels. To remove these unwanted components, it is usual practice to place a honeycomb

directly upstream of the damping screens. For tunnels of the size being con- sidered here, satisfactory performance has been obtained with a cell diameter

of 2 inches (50.8 mm) or less and a cell length of about 8 times the cell

diameter [22].

When a honeycomb is used as a flow conditioner, it is necessary to provide

sufficient length in the downstream portion of the settling chamber to allow

the turbulence generated by the honeycomb to decay before the flow enters the

contraction. Alternatively, one or more damping screens can be placed downstream

of the honeycomb. Prior to the work of Loehrke and Nagib [23], the mechanisms

by which honeycombs suppress incoming turbulence and generate turbulence in the

exiting flow were not well understood. As a result, the benefits expected of

honeycomb installations were not always realized and in certain cases the flow

quality was actually reduced by the installation of a honeycomb [24].

The work reported in reference 23 indicates that the suppression of turbulence

in the flow entering a honeycomb is due in large part to the inhibition of the

lateral components of fluctuating velocity. Flow visualization studies show

that most of the suppression occurs within the first 5 to 10 diameters along

the honeycomb cells and that the predominant mechanism by which turbulence is

generated in the flow exiting the honeycomb is shear-layer instability. By

placing a fine-mesh damping screen (K = 0.86) within 5 cell diameters of the

downstream face of the honeycomb, loehrke and Nagib showed that the shear

layers in the exiting flow can be modified, leading to very rapid decay of

turbulence and reducing the intensity to approximately 2 percent at a downrstream distance of from 30 to 40 cell diameters.

3.6.3 Arrangement of Flow Conditioners

The recommended arrangement of the settling chamber is shown in figure 6. A

2 ft (610 mm) space is provided at the settling chamber inlet to gain access to

the honeycomb and to allow for cleaning of the back side of the last diffusing

screen in the wide-angle diffuser. This space also allows for some lateral

adjustment of the flow before entering the honeycomb. Because they are inexpensive and readily available, 2 in (50.8 mm) O.D. cardboard tubes with a length of 10 diameters or 20 in (508 mm) are recommended for construction of

the honeycomb.

Following the recommendation of reference 23, a damping screen is installed

approximately 4 tube diameters or 8 in (203 mm) from the downstream face of the

honeycomb. The choice of K value for the damping screen does not appear to be

critical. However, a screen with K = 0.86 was used in the studies reported in

reference 23 and this is close to the head-loss coefficient K = 0.70 selected

for the diffusing screens. With this combination of honeycomb and damping

screen, the intensity of the streamwise component of turbulence should be

approximately 2 percent at 5 to 6 ft (1.52 to 1.83 m) downstream of the screen.

16
Additional reduction of turbulence will be achieved in the contraction and

this is discussed later.

The loss coefficients for the honeycomb and damping screen are shown in

reference 23 to be directly additive. Reference 24 suggests K = 0.30 for

tubular honeycomb of the type used here. This value corresponds to a length/

diameter ratio of 6 and, therefore, K = 0.5 would be a reasonable choice for a length/diameter ratio of 10. Thus the combined loss coefficient for the honeycomb and damping screen would be K = 0.5 + 0.7 = 1.2.

In order to assess the effectiveness of the honeycomb/screen combination in the

settling chamber, the level of turbulence directly ahead of the honeycomb must

be known. Based on information presented in reference 25, a turbulence intensity of about 12 percent toward the end of a well-designed return circuit is

considered reasonable. For the tunnel being considered here, the use of a relatively rough circular duct and a wide-angle diffuser in the return circuit

can be expected to result in a higher level of turbulence at the fourth corner,

say 20 percent. With a loss coefficient of K = 0.70 for each of the three

diffusing screens, the reduction in the r.m.s. level of the streamwise fluctuations as the flow passes through the second wide-angle diffuser will be about

55 percent, according to equation (19). Accounting for the change in mean

velocity between the fourth corner and the settling chamber, the intensity of

the streamwise component of turbulence just upstream of the honeycomb will be

about 18 percent, based on the local flow velocity. This estimate does not

account for natural decay or production of turbulence by the second wide-angle

diffuser. Nevertheless, the honeycomb/screen combination appears to be far more efficient than are multiple damping screens or a honeycomb and screens without proper attention given to the honeycomb-screen spacing. Assuming that

a series of damping screens with K = 0.70 were substituted for the honeycomb/

screen combination, 8 screens would be required to achieve the same reduction

in turbulence intensity. The combined head-loss coefficient would be 8 x 0.70

5.6 vs. 1.2 for the honeycomb/screen combination.

3.7 EFFECT OF THE CONTRACTION

The contraction section for the tunnel arrangement shown in figure 1 has a contraction ratio (inlet area/outlet area) of 3.2:1 and an aspect ratio (width

of section/height of section) of 1.33:1 at the inlet and 1.67 at the outlet.

Although the details of the contraction design are addressed later, the effects

of the contraction on the flow are almost entirely dependent upon the contraction ratio. Therefore, these effects are addressed here in the context of

flow conditioning.

3.7.1 Nonuniformity of the Mean Flow

An important function of a wind tunnel contraction is the generation of uniform

mean flow at the entrance to the test section. Assuming that the flow in the

contraction is incompressible, inviscid and steady, it can be shown [26] that

the improvement in flow uniformity on passing through the contraction is given

by

17
1

( 21 ) < UA2 - V/UA2 U

1 + UA1 A1

(UA1 - Ui)/Uai u2 + UA2 UA2 CR

In this expression is the area-averaged mean velocity, U is the local mean

velocity, Cr is the contraction ratio and the subscripts 1 and 2 denote the

inlet and outlet sections, respectively.

For a contraction ratio of 3.2:1 the attenuation factor for nonuniformity of

mean velocity is 0.10. Thus the local nonuniformity downstream of the

honeycomb/damping screen should not deviate from the mean velocity by more

than 10 percent if the target uniformity of + 1 percent in the test section is

to be met. Of course this excludes the boundary layer region at the walls of

the contraction. With reasonable care in the installation of the diffusing

screens, honeycomb and damping screen, there should be no difficulty in satisfying the 10 percent limitation on velocity deviation directly upstream of the

contraction.

3.7.2 Reduction of Turbulence

The effects of a contraction on free-stream turbulence using the simple concept

of preferential stretching and shortening of cylindrical vortex filaments and

using the concept of rapid distortion (sometimes referred to as the linear theory) are summarized in reference 26. Applying the concept of preferential stretching

and shortening, the intensities of the three components of turbulence, u, v and

w, are attenuated by an axisymmetric contraction with Cr = 3.2:1 as follows:

1

? 1 = -4 = 0.10 ( 22 ) I ul C

2R

I v2I w2

= 0.56

vl

1 wli R

(23)

Rapid distortion theory leads to the following relationship between pre- and

post-contraction turbulence intensities with Cr = 3.2.

0.17

^v2 _ ^w2

Ivl I wl

0.49

(24)

(25)

18
Experimental studies carried out by Ramjee and Hussain [27] using axisymmetric

contractions with contraction ratios ranging from 11:1 to 100:1 indicate that

the rapid distortion theory overestimates the attenuation of the intensity of

the streamwise component and underestimates that of the transverse components.

For contraction ratios of 11:1 or less, data presented in reference 27 suggest

that I v2/-*-u2 ~ Therefore, use of equation (25) will provide a conservative

estimate of the turbulence intensity at the test section entrance. Assuming

that the honeycomb/screen combination installed in the settling chamber will

perform as expected (I u i = I v i = 0.02), the turbulence intensity at the entrance

to the test section should be approximately 1 percent (I v2 = 0.49 Iv l)* Further

reduction of the turbulence intensity will require additional damping screens

in the settling chamber.

3.8 DESIGN OF THE CONTRACTION

Factors that influence the choice of contraction geometry include the degree to

which spatial irregularities in the velocity distribution are to be reduced,

the intensity of turbulence expected at the entrance to the test section, and

the effect of the contraction ratio on tunnel power consumption. If highly

uniform flow of very low turbulence (/^Z/U < 0.001) is desired, contraction ratios

of 9 or greater are usually required. In the case of closed-circuit tunnels

the use of large contraction ratios results in relatively low speeds in the

return section and a corresponding reduction in losses, particularly at the

third and fourth corners and at the damping screens. For an open-circuit

tunnel the greatest reduction in losses occurs at the damping screens and, to

a lesser extent, at the tunnel inlet. Factors that argue against a large

contraction ratio are the increased floor area and headroom required for the

third and fourth corners and for the settling chamber, the increased length of

the contraction (which is usually at the expense of useable test section

length), and the cost of construction.

Available floor space in this case limits a closed-circuit design to an

over/under arrangement as shown in figure 1. The dimensions of the contraction

inlet could be increased beyond the 16 x 12 ft (4.877 x 3.658 m) section indicated,

but this would have to be at the expense of clear working space along the

sides of the settling chamber (already considered to be the minimum acceptable)

and headroom available for installation or replacement of the diffusing

and damping screens. Since the arrangement of the return duct, the third and

fourth corners, and the wide-angle diffusers is dictated by factors other than

the contraction ratio, a contraction ratio larger than the value of 3.2 used

here would not lead to a significant reduction in power requirements. If a lower level of turbulence at the entrance to the test section is desired, this can better be achieved with the installation of additional damping screens

in the settling chamber.

For the reasons just described, the contraction ratio is by far the most important

parameter in the design of a wind tunnel contraction. Once this ratio has been

selected, the wall profiles and the overall length of the contraction must be

established. Ideally, the contraction should operate over the design speed

19
range without flow separation, the boundary layer should be relatively thin,

and the exit velocity should be highly uniform. These features should be

achieved within the shortest possible length.

As Bradshaw and Pankhurst [25] have noted, there is no method, either in two or three dimensions, that can be considered entirely satisfactory for contraction

design. Although various analytical techniques have been used in the design of

axisymmetric and two-dimensional contractions, many low-speed tunnels of recent

construction use contractions that either are adaptations of previous "good"

designs that have given satisfactory performance, or are based upon systematic

studies using scale models. A major problem in contraction design is that for

any contraction of finite length, the flow along the wall must pass through

points of minimum and maximum velocity. Thus there is a region near each end

of the contraction where the pressure gradient along the wall is adverse and

where, under certain conditions, separation of the boundary layer can occur. Flow separation is a distinct possibility in the case of small contraction

ratios where the flow accelerations may be too small to overcome locally

adverse pressure gradients, and in the case of two- and three-dimensional

contractions where secondary flows develop in the boundary layer at the

corners

.

The contraction profile considered here is an adaptation of one described

in reference 28. Tests carried out in both model and full scale indicate a highly uniform exit velocity and the absence of boundary layer separation. The

wall profile, derived from a 4th order polynomial, has the form

4 3

Y = -3h(l- f) + 4h( 1- f) (26)

Li Li where X, Y, h and L are as defined in figure 7.

Based on overall dimensions of contractions that have demonstrated good

performance and have contraction ratios close to the ratio being considered

here [4], a contraction length of 14 ft (4.267 m) was selected. To simplify

fabrication of the contraction and to make it easier to install corner fillets

if their use is desired at some future time, the same profile is used for all

four walls. Dimensionless coordinates of the wall profile, the cross sectional

area, the local pressure coefficient based on average axial velocity and

referenced to the static pressure at the outlet, and the pressure gradient

along the axis of the contraction are listed in table 1. The local pressure

coefficient and the pressure gradient are plotted in figure 8. Also plotted in

figure 8 are C

p

and its gradient for two wind tunnel contractions of recent

construction [29,30] whose dimensions are reasonably close to those of the

contraction being considered here. Note that the pressure gradients for these

two tunnels are approximate, having been estimated from the distribution of C

p since the exact form of the wall profiles is not known. Overall dimensions of

the three contractions are listed in table 2. It can be seen from figure 8

that the pressure gradient in the "core flow” is favorable for all three

contractions, the maximum absolute value for the proposed contraction being

approximately 90 percent of the other two.

20
While such a comparison may be reasonably accurate for the flow in the core of

the contraction, it is not possible to extend this comparison to the flow along

the walls or, more importantly, to the flow in the corner regions where boundary

layer separation will occur first.

To assess the potential for boundary layer separation, reference is made to the

work of Morel on axisymmetric contractions [31] and on two-dimensional wind

tunnel contractions [32]. The approach taken by Morel was to solve the Euler

equations of motion using a finite-difference procedure. The boundary layer

displacement thickness was assumed to be much smaller than the transverse

dimensions of the contraction so that the flow could be treated as potential

flow. The boundary conditions were the wall profiles consisting of two smoothly

matched arcs with apex at the ends of the contraction and a finite length of

straight duct at each end of the contraction where the velocity profile approaches

uniform flow. Using the results of this analysis, Morel established a procedure

for selecting a shape parameter and contraction length that just avoids boundary

layer separation and minimizes the boundary layer thickness. Based on the

analysis of more then 100 configurations with power-law wall shapes of order

ranging from 2 to 5, Morel concluded that two smoothly joined cubic arcs satisfy

the separation criterion for an axisymmetric contraction with relatively short

length. The analysis of two-dimensional contractions reported in reference 32 was limited to wall profiles consisting of two cubic arcs.

The three-dimensional contraction used in the tunnel arrangement shown in figure

1 can be considered to be bounded by the two-dimensional and axisymmetric cases analyzed by Morel. If the contraction is assumed to be two-dimensional, the

contraction ratio is equal to the ratio of the inlet height to outlet height, or 2.0. The inflection point of the wall profile, which is located at a distance

of L/3 from the inlet plane, can be taken as the equivalent of the match point

for Morel’s cubic arcs. An additional parameter required to establish Morel's

separation criterion is the distance from the virtual origin of the boundary

layer to the inlet plane of the contraction which is taken to be the distance

to the last damping screen in the settling chamber; approximately 4 ft (1.22 m)

in this case. The wall pressure coefficient associated with separation is

Cpi = 0.41 and the expected value is Cp^ = 0.31 where

cpl = 1 - oyv) 2 (27)

In this definition of Cp£ , U-^ is the mean velocity sufficiently far upstream

of the inlet for the flow to be considered uniform and V is the local velocity

just outside of the boundary layer.

For the axisymmetric case the contraction diameters are taken to be 4 times the

hydraulic radius at the inlet and outlet planes of the three-dimensional contraction. The corresponding contraction ratio is 3.34 and the critical value

of Cp^ is again 0.41. This compares with an expected value of Cp^ = 0.28.

Since these pressure coefficients are based on wall profiles consisting of cubit

arcs, they are not directly applicable to the 4th order profile being cons id. r. I here. For an axisymmetric contraction with contraction ratio of 9, Morel

investigated the effect of different power-law wall shapes on Cp[ and on the

21
pressure gradient term used to establish the criterion for flow separation. The

value of Cpi for a 4th order power law curve is approximately 1.35 times that

for a cubic arc while the pressure gradient term is essentially unchanged. The

effect for a contraction ratio of about 3 probably would be less. Nevertheless,

the effects of secondary flow in the boundary layer adjacent to the corners of

the three-dimensional contraction are cause for concern and the final design

should allow for the possible installation of corner fillets to reduce the

tendency for secondary flows to develop in the corners of the contraction.

With regard to the potential for boundary layer separation near the end of the

contraction, Morel's results for an axisymmetric contraction indicate the local

pressure coefficient at the wall is not strongly dependent on the choice of

power-law wall profile [31]. Since the length selected for the contraction

being considered in this design is slightly longer than the "optimum" length

for a wall profile consisting of matched cubic arcs, separation near the end

of the contraction is not likely to be a controlling factor.

3.9 ESTIMATION OF ENERGY RATIO

The energy ratio for a wind tunnel is defined as the ratio of the flux of

kinetic energy at the test section entrance to the rate of work done by the

drive section on the fluid. Assuming uniform flow at the test section entrance,

the energy ratio can be expressed as

where Aq and UQ are the area and velocity, respectively, at the test section

entrance and AP t is the total head loss along the path length of the tunnel.

The total head loss, APt , is equal to the sum of the losses due to the individual

tunnel components and may be expressed as

1/2 pU3 A

o o (28) A0 U0 APt

APt = ^ 1/2 pU? = Koi 1/2 pU;r2 (29) o

and thus

( 30 )

22
where is the appropriate reference velocity for calculating the head loss

through the tunnel component. For a tunnel with n discrete components that

contribute to the total head loss, we have

n

Ap

t = 2 K

q

. 1/2 pUQ

2 (31)

i=l

and

E.R. = — (32) n

Z Koi

i=l

In most cases is related to the local Reynolds number with K-^ increasing

with decreasing Re

. For purposes of estimating power requirements the losses

are based on the associated with maximum Re where it is usual to observe

only a weak dependence of K-^ on Rg .

Table 3 lists the tunnel components, the estimated values of , the source of

these estimates and the corresponding values of K0 ^. For purposes of establishing the maximum power requirements, the test section is assumed to contain a bluff body (circular cylinder) spanning the test section from floor to ceiling

and having a characteristic dimension of 0.5 ft (152 mm). Also listed in

table 3 is the percentage contribution of each component to the total head

loss. The estimated energy ratio for the closed-circuit tunnel is 0.53.

The energy ratio is an indicator of tunnel efficiency and a well-designed tunnel

is characterized by both a high energy ratio and satisfactory performance

characteristics. Aeronautical tunnels with a long diffuser in the return

circuit and operating without screens or other flow conditioning devices may

have energy ratios of 5 or greater [24].

3.10 PRESSURE PROFILES

Once the head losses have been estimated, it is possible to calculate the

profiles of total pressure and static pressure around the tunnel circuit.

The total pressure profile reflects the head losses along the tunnel circuit

while the static pressure profile is used to obtain the design loads for the

various tunnel components. The last two columns of table 3 list the total

pressure and static pressure, respectively, for an operating speed of 80 fps

(24.38 m/s). The corresponding profiles are plotted in figure 9.

Unless provisions are made for venting a closed-circuit tunnel, the static

pressure will adjust to the leakage that is almost always present in construction joints, screen slots, etc. For a "sealed" tunnel, the static pressure in

regions of high velocity will be lower than the outside ambient pressure.

This usually means that the test section will operate at a static pressure

substantially below the outside ambient pressure with the result that any

23
leaks in the test section will act as inward-directed jets that can disrupt

the growth of boundary layers and alter the base pressure for models undergoing

tests. This problem can be reduced by exercising great care in the design and

construction of the test section, but the most effective step is to intentionally vent the test section to the outside ambient pressure. This is usually

done at the downstream end of the test section so that most of the test section

operates just slightly above ambient pressure. Sealing of fixtures such as turntables, cable entries, etc. then becomes relatively unimportant. The

penalty for venting the test section is an increase in static pressure throughout the tunnel and this can lead to significant increases in design loads for

components in regions of very low speed, e.g. the third and fourth corners and

the settling chamber. The static pressure profile plotted in figure 9 assumes

venting just downstream of the first corner.

3.11 FAN SELECTION

Knowing the energy ratio of the tunnel, the relationship between the flow rate

and the head loss can be obtained. This relationship is plotted as the tunnel

load characteristic in figure 10. Note that the tunnel load characteristic at

very low flow rates is underestimated because of the dependence of loss

coefficients on the Reynolds number. This will influence the required fan

speed at low rates of flow but will not affect the fan size requirements.

Characteristic curves for a typical axial flow fan with an 8 ft (2.438 m)

diameter casing are also plotted on figure 10. The operating point corresponding to maximum fan efficiency (approximately 76 percent for this fan) is

indicated on each of the fan characteristic curves.

To ensure stable tunnel speed, the fan characteristic curves should have

relatively steep slopes over the operating range and should intersect the tunnel

load characteristic curve at angles approaching 90 degrees. From figure 10 it

is seen that stability may present a problem for tunnel speeds of less than

about 30 fps (9.14 ra/s). Methods that have been used to overcome this problem

include; (1) equipping the tunnel with a bypass section so that the flow rate

through the fan remains high for all tunnel speeds, (2) using a fan with con- tinuously variable, remote-controlled pitch setting or (3) using a fan with

incremental pitch settings (made with the fan stopped). The first two methods,

while more convenient, are relatively expensive. The third method offers

substantial improvement in performance at little additional cost. The dashed

line in figure 10 is a fan characteristic for the case of selectable-pitch

blades set at a relatively flat angle of attack.

For a maximum tunnel speed of 80 fps (24.38 m/s), the corresponding flow rate is

288,000 cfm (136 m-^/s) and the tunnel load characteristic indicates a head loss

of approximately 2.65 in (67.3 mm) of water. The required fan speed for a typical 8 ft (2.438 m) diameter axial flow fan is about 690 RPM and the corre- sponding speed of the blade tips is 289 fps (88.09 m/s). This is in general

agreement with the maximum tip speeds for the low-speed tunnels described in

reference 4 and easily satisfies the recommendation of reference 25 that tip

speed not be allowed to exceed half the speed of sound. Based on the tunnel

load characteristic and a maximum operating speed of 80 fps (24.38 m/s), an

24
axial flow fan with an 8 ft (2.438 m) casing diameter will be adequate. To ensure stable operation at low tunnel speeds, U0 < 30 fps (9.14 m/s), the fan

should be equipped with selectable-pitch blades.

3.12 POWER REQUIREMENTS

The power required at the fan shaft is plotted in figure 11. This curve is

based on an overall fan efficiency of about 72 percent and does not account for

uncertainties in the tunnel energy ratio, motor efficiency or mechanical losses

between the motor and fan. Therefore, motor selection should be based on a power requirement approximately 10 percent greater than is indicated by

figure 11 or about 180 hp (134 kW). A motor of this size can be housed within

the fan nacelle with direct coupling to the fan hub.

For motors larger than 150 hp (112 kW), the size selection offered by most

manufacturers is limited to increments of 50 hp (37 kW) . Allowing for a 10 percent margin on required power, a 150 hp (112 kW) drive motor will be

adequate for tunnel speeds of up to 7 5 fps (22.86 m/s). The consequences of

setting the maximum operating speed below the target value of 80 fps (24.38

m/s) are discussed later in the context of an open-circuit tunnel.

3.13 MOTOR AND CONTROLLER

The size of motor being considered here and the fact that it will be coupled

directly with the fan preclude the use of a number of mechanical and electro/

mechanical devices that are available for speed control. Of the five means of

electrical speed control described in reference 38, it appears that frequency

regulation of an induction motor or voltage regulation of a direct current

motor are the only practical alternatives.

The principal advantages of frequency regulation are the high reliability and

low maintenance inherent in an induction motor and the fact that the torque/

speed characteristic changes only slightly as the supply frequency changes.

The disadvantages of this method of speed regulation are the relatively high

cost of the controller and the limited range of speed regulation. The "turndown ratio” or factor by which the speed can be changed is limited to about

10:1 so that the minimum tunnel speed will be approximately 10 percent of the

maximum speed for a given pitch setting. This may or may not be a problem,

depending upon the fan characteristics. With direct coupling to the fan and a closed-circuit tunnel, there are no other feasible methods for achieving very

low tunnel speeds if an induction motor is used to drive the fan.

Compared with a.c. motors, speed control using a d.c. motor is relatively

simple. Its main advantages are that speed control is available down to zero speed and this greatly simplifies the problem of starting a system with high

inertia. However, maintenance costs are high compared with induction motors.

The development of compact solid-state rectifiers has increased the reliability

and reduced the cost of d.c. power supplies with the result that this method of

speed control is widely used in wind tunnels requiring drive motors larger than

about 50 hp (37 kW.)

25
3 . 14 SUMMARY

The performance requirements for a low-speed boundary layer wind tunnel can be

met with a closed-circuit design employing an over/under arrangement and a contraction ratio of 3.2:1. Recommended test section dimensions are W = 10 ft

(3.048 m), H = 6 ft (1.829) and L = 55 ft (16.764 m), measured to the turntable.

Because of space limitations, it is necessary to place severe restrictions on

the transverse dimensions of the return section with a correspondingly low

energy ratio (E.R. = 0.53).

An axial flow fan with a diameter of 8 ft (2.438 m) is recommended. Both

variable pitch and variable speed will be required to ensure stable operation

over the specified range of speed. Pitch control under load is not necessary

and this feature can be provided by a fan hub with manually adjustable blades.

The estimated maximum power demand is based on a fan efficiency of 72 percent,

approximately 4 points below peak efficiency for the fan considered here. It

is anticipated that an operating efficiency in the range 75-80 percent can be

achieved with proper choice of a standard production model fan. Maximum power

demand will be approximately 180 hp (134 kW) and this can be reduced to 150 hp

(112 kW) if the maximum tunnel speed is reduced to 75 fps (22.86 m/s).

Either an induction motor with variable frequency control or a d.c. motor with

variable voltage control can be used for speed regulation. The d.c. drive

system is expected to have a lower initial cost, but a higher maintenance

cost

.

4. OPEN-CIRCUIT TUNNEL DESIGN

4.1 GENERAL

The open-circuit tunnel arrangement shown in figure 12 should be considered as an alternative to the closed-circuit design if it is determined that a thermal

modeling capability is not essential to the conduct of CBT research activities.

The major advantages of this arrangement are its lower initial cost, easier

access to the turntable near the end of the test section, and most important,

operation of the test section at a static pressure just slightly above outside

ambient pressure. Major disadvantages are a greater power demand for the same tunnel speed and a less comfortable laboratory working environment.

The open-circuit design makes use of the same settling chamber and contraction

described earlier for the closed-circuit arrangement. Because of different

inlet conditions and a different type of fan, a redesign of the wide-angle

diffuser is required. The test section, although slightly shorter, has many of

the same features proposed for the closed-circuit scheme.

4.2 WIDE-ANGLE DIFFUSER

The wide-angle diffuser for the open-circuit arrangement is shown in figure 13.

Since the diffuser inlet geometry must be compatible with the discharge side

26
of the fan, the inlet dimensions indicated in figure 13 are based upon dimensions

of a typical fan having the required capacity.

For reasons discussed later, the tunnel design assumes a centrifugal fan. An

undesirable feature of this type of fan is the highly nonuniform distribution

of velocity at the fan outlet. Not only is this distribution highly nonuniform,

but the distribution changes significantly with fan speed. For this reason a plenum and a diffusing screen are incorporated with the diffuser inlet. As has

been discussed elsewhere in this report, a screen with head loss coefficient

K = 2.8 has been shown to be most efficient in promoting flow uniformity.

Splitter plates or vanes have been incorporated in the diffuser design to

prevent transitory separation and stalling. The reason for selecting splitter

plates over diffusing screens is that some provision must be made for silencing

the fan and although there is room in the plenum for splitter plates, they

would interfere with the redistribution of flow that must occur within the

plenum.

The divergence angle for the wide-angle diffuser is 18 degrees in the vertical

plane and 24 degrees in the horizontal plane. A 3 x 3 cell array formed by the

splitter plates would reduce the individual divergence angles to a maximum of 8

degrees which is considered to be an upper limit for stable operation of the

diffuser. However, there are two arguments for adopting a 4 x 4 cell array.

First, some "fine tuning" of the splitter plates will be required to establish

equal flow rates in the individual cells and this can best be accomplished with

vertical and horizontal splitters located in the planes of symmetry of the

diffuser. Second, the efficiency of the splitter plates as acoustic silencers

will increase with increasing length to width ratio of the cells.

The head loss coefficient for the diffuser is estimated by treating the

individual cells as equivalent conical diffusers with an outlet to inlet area

ratio of 1.55 and a ratio of length to inlet radius of 5.33. According to the

flow regime boundaries defined in reference 36, a conical diffuser with these

parameters can be expected to operate with attached flow throughout the length

of the diffuser. Note that the splitter plates stop short of the end of the

diffuser to allow for redistribution of the flow directly ahead of the second

diffusing screen.

The pressure recovery coefficient and the total-pressure loss coefficient are obtained from reference 36 based on the assumption that the settling chamber,

honeycomb and screens function effectively as a tailpipe for the 16 diffuser

cells. The corresponding values of these coefficients are Cp r = 0.50 and K L = 0.13. In order to suppress the wakes created by the splitter plates, the

head loss coefficient of the downstream diffusing screen is set at K = 2.8.

Because of the high loss due to the diffusing screens at each end of the wideangle diffuser, uncertainties as to the value of the total-loss coefficient,

for the diffuser are relatively unimportant.

27
4.3

TEST SECTION

The test section has the same general features as that for the closed-circuit

arrangement. Because of the space required for the fan, plenum and free

discharge at the end of the tunnel, the overall length of the test section is

limited to 60 ft (18.288 m). A distance of 10 ft (3.048 m) has been provided

between the center of the main turntable and the end of the test section for

wake development and decay. This distance appears to be sufficient based on a number of boundary layer tunnels described in reference 4. If additional length

is required, as may be the case for very large bluff models, a temporary

extension of the floor plane can be installed.

4.4

ESTIMATION OF ENERGY RATIO

Head loss coefficients are referenced to the test-section dynamic pressure as was done for the case of the closed-circuit tunnel. These coefficients are listed in table 4 and the energy ratio is estimated to be 0.39. Compared with

the energy ratio of 0.53 for the closed-circuit tunnel, the open-circuit

arrangement is seen to be substantially less efficient. The major losses are due to the first diffusing screen at the juncture of the plenum and the

wide-angle diffuser (35 percent) and the loss due to free discharge at the

end of the test section (40 percent).

Depending upon the flow nonuniformity at the discharge side of the fan, it is

possible that with proper choice of fan the head loss coefficient for the first

diffusing screen can be reduced without adversely affecting the performance of

the wide-angle diffuser. However there are no practical means available for

reducing the loss coefficient at the end of the test section. A simple diffuser

or a diffuser with turning vanes would defeat the purpose of the proposed arrangement, i.e. ease of access to the turntable and a test section static pressure

close to ambient pressure. In addition, the space required for an efficient

diffuser would reduce the length of the test section, already considered to be

the minimum acceptable. With the proposed arrangement, a clearance of approximately 20 ft (6.10 m) has been provided between the tunnel outlet and the

laboratory wall. Measurements reported in reference 30 suggest that instability

of the discharge jet and a corresponding unsteadiness in the tunnel speed can

result if this clearance is too small.

4.5

PRESSURE PROFILES

The total pressure and static pressure for the open-circuit tunnel operating at

a speed of 80 fps (24.38 m/ s) are listed in table 4 and are plotted in figure 14.

The pronounced losses in total pressure due to the first diffusing screen

and the free discharge are readily apparent. Exact distributions of the total

and static pressure ahead of and through the fan will depend upon the shape of

the fan inlet and the fan characteristics.

4.6

FAN SELECTION

As with the closed-circuit tunnel, the major criterion for choice of fan type

is high capacity at relatively low total pressure head and this narrows the

28
selection to the axial flow and centrifugal types. If available space were

not a consideration, the axial flow type would be the obvious choice because

the motor and fan can be directly coupled within the fan casing and the fan discharge can be conditioned more efficiently than can that of a centrifugal fan.

However, available space is an important limitation and a substantial reduction

in overall length of the tunnel can be realized by using a centrifugal fan

with a double inlet. In order to simplify the transition to the settling

chamber, it is desirable that the dimensions of the fan outlet be as close as possible to those of the settling chamber.

The tunnel load characteristic is plotted in figure 15 along with characteristic

curves for a typical double-wheel, double-inlet, backward-b laded centrifugal

fan with wheel diameter of 98 in (2.489 m). Operating points for maximum fan

efficiency (82 percent) are indicated on the characteristic curves. For this

particular fan, the operating efficiency will be approximately 73 percent. At

maximum tunnel speed the fan speed will be about 360 rpm and the speed of the

blade tips will be 154 fps (46.94 m/s). This is well below the limit of half

the speed of sound recommended for low-speed tunnel applications [25].

The discussion of system stability for axial flow fans (Sec. 3.11) applies to

centrifugal fans as well. Since the configuration of the fan wheel cannot be

changed as can the blade pitch for an axial flow fan, it is usual to install

variable-pitch vanes on the inlet bell of the fan. By adjusting the inlet vanes the flow rate can be regulated down to about 25 percent of maximum with a constant speed drive. In order to satisfy the requirement for continously

variable tunnel speeds over the range of 1 to 80 fps (0.30 to 24.38 m/s), a variable speed drive will be required.

4.7 POWER REQUIREMENTS

Fan input power is plotted against tunnel speed in figure 16. The power

requirement is based on a fan efficiency of 73 percent and does not account for

uncertainties in estimating the energy ratio of the tunnel. Allowing a 10

percent margin for this uncertainty and for mechanical losses, the power

requirement is approximately 250 hp (187 kW) . For normal operation at speeds

of about 40 fps (12.19 m/ s) the power demand will be approximately 30 hp

(22 kW).

By reducing the motor size to 200 hp (149 kW) , the maximum tunnel operating

speed will be approximately 75 fps (22.86 m/s), equal to that of the closedcircuit tunnel equipped with a 150 hp (112 kW) motor. The corresponding

reduction in initial cost of the motor and controller is estimated to be 20 to

25 percent. Recalling that the requirement for maximum speed is based on the

need to achieve critical Reynolds numbers without excessive tunnel blockage , increasing the body diameter by a factor of 8/7.5 would offset the effect of

speed reduction. The corresponding solid blockage would be 5.3 vs 5.0 percent . This slight increase in solid blockage is considered acceptable in view of the

cost reduction that can be realized with a 200 hp (149 kW) drive motor in tin- open-circuit tunnel or a 150 hp (112 kW) drive motor using a closed-circuit

arrangement

.

29
4.8 MOTOR AND CONTROLLER

Requirements for the motor and speed control system are much the same as those

for the closed-circuit tunnel. A centrifugal fan allows the choice of direct

coupling or belt drive, each arrangement having certain advantages and disadvantages. Because of its higher inertia, the centrifugal fan will require

more attention to the design of the starting and speed control system.

4.9 SUMMARY

For applications that do not require a thermal modeling capability, an opencircuit tunnel will be adequate. A tunnel of this type can incorporate the same settling chamber, contraction and test section that have been designed for

the closed-circuit arrangement. Because of minimum end clearance required for

the tunnel discharge, the test section length must be reduced to 50 ft

(15.240 m) as compared with 55 ft (16.764 m) for the closed-circuit arrangement.

A backward-b laded centrifugal fan with a double inlet is recommended for the

open-circuit design. Because the flow on the discharge side of this type of

fan is highly nonuniform, flow conditioning requires the use of diffusing

screens with high head loss coefficients. This, combined with the discharge

loss at the end of the tunnel, results in a relatively low energy ratio (E.R.

= 0.39). For a fan operating efficiency of 73 percent the maximum power demand

is approximately 250 hp (187 kW) . If the maximum tunnel speed is reduced to

75 fps (22.86 m/s), a 200 hp (149 kW) motor will be adequate.

Speed regulation will require both inlet vanes and a variable speed drive to ensure stable tunnel operation at all speeds. Either an induction motor with

a variable frequency controller or a d.c motor with a variable voltage controller

can be used.

5. CONCLUSIONS

Based on the anticipated use of the boundary layer wind tunnel in support of

research activities within the Center for Building Technology, it is concluded

that tunnel operating speeds over the range of 1 to 80 fps (0.30 to 24.38 m/s)

will be sufficient for a majority of applications. A turbulence intensity of

approximately 0.5 percent and a maximum nonuniformity of flow on entry to the

test section of + 1 percent can be achieved using a simple honeycomb/screen

combination and a contraction ratio of 3.2:1.

A test section width of 10 ft (3.048 m) and a height of 6 ft (1.829 m) will

be adequate for 1:500 scale simulations of the atmospheric boundary layer and

for achieving critical Reynolds numbers at or near the maximum operating speed

of 80 fps (24.38 m/s). Available space limits the test section width and

length. Depending on the tunnel configuration selected, test section length

can range from 50 to 55 ft (15.240 to 16.764 m).

Although provisions for thermal modeling have not been considered specifically

in this preliminary design effort, it is clear that such a capability will

require a closed-circuit tunnel as shown in figure 1. Because of limitations

30
on laboratory space, an over/under arrangement is the only alternative and

this will involve relatively high construction costs for the return leg of

the tunnel.

If a thermal modeling capability is not considered essential to CBT research

activities, the requirements for speed range and size of test section can be

met with an open-circuit design as shown in figure 12. However, the power

demand will be higher and the laboratory working environment will be less

comfortable.

An axial flow fan with an 8 f t (2.438 m) diameter has been assumed in the design

of the closed-circuit tunnel. Both variable pitch and variable speed control

will be required for stable operation over the operating range of the tunnel.

It is not necessary to vary the pitch while the fan is operating. For the opencircuit design a centrifugal fan with double inlet and a wheel diameter of

98 in (2.489 m) will be adequate. Adjustable inlet vanes and a variable-speed

drive will be required for tunnel speed regulation.

Either an induction motor with a variable frequency controller or a d.c. motor

with a variable voltage controller can satisfy the requirements for regulation

of fan speed. A d.c. motor and controller are expected to have a lower initial

cost, but the maintenance costs are expected to be higher than those for an induction motor with variable frequency control. Power currently available in

the wind tunnel construction area consists of 480 volt 3-phase service and

120/240 volt d.c. service. The d.c. service is limited to 420 amperes which

is adequate for normal tunnel operation but is insufficient to meet the maximum

power demand. Therefore, a d.c. drive system must include a rectifier which

is consistent with motor size. Limits on 3-phase service do not pose any

problems

.

Reducing the maximum tunnel speed to 75 fps (22.86 m/s) will reduce the maximum

power demand from 180 hp (134 kW) to 150 hp (112 kW) in the case of the closedcircuit tunnel. Depending on the motor/controller price structure, this reduction of maximum power demand may or may not be significant. A larger reduction

can be realized for the open-circuit design; 200 hp (149 kW) vs 250 hp

(187 kW) and it will likely prove cost effective to limit the power demand to

200 hp (149 kW) . The adverse effect of limiting tunnel speed to 7 5 fps

(22.86 m/s) would be to increase the solid blockage from 5.0 to 5.3 percent

in order to achieve critical Reynolds numbers using a smooth circular cylinder

as a test body. In view of the associated cost reduction, this slight increase

in solid blockage is considered acceptable.

31
6.

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32
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22. Stathopoulis , T., "Design and Fabrication of a Wind Tunnel for Building

Aerodynamics," Journal of Wind Engineering and Industrial Aerodynamics , Vol. 16, No. 2, April, 1984, pp. 361-376.

23. Loehrke, R. I. and Nagib, H. M., "Control of Free-Stream Turbulence by

Means of Honeycombs: A Balance Between Suppression and Generation,"

Transactions fo the ASME, Journal of Fluids Engineering , Vol. 98, No. 3,

September 1976, pp. 342-353.

24. Pope, A. and Harper, J. L. , Low-Speed Wind Tunnel Testing , John Wiley & Sons, Inc., New York, NY, 1966.

25. Bradshaw, P. and Pankhurst, R. C., "The Design of Low-Speed Wind Tunnels,''

Progress in Aeronautical Sciences, Vol. 5, Pergamon Press, 1964, pp.

33
26. Hussain, A. and Ramjee, V., "Effects of the Axisymmetric Contraction Shape on Incompressible Turbulent Flow," Transactions of the ASME, Journal of

Fluids Engineering , Vol. 98, No. 1, March, 1976, pp. 58-69.

27. Ramjee, V. and Hussain, A., "Influence of the Axisymmetric Contraction

Ratio on Free-Stream Turbulence," Transactions of the ASME, Journal of

Fluids Engineering , Vol. 9 8, No. 3, September, 197 6, pp. 506-515.

28. Fujii, K. , Ueda, H. and Sato, S., "On the Basic Design and Construction of an Industrial Aerodynamics Wind Tunnel," Journal of Wind Engineering,

No. 13, July, 1982, pp. 47-66.

29. Snyder, W. H., "The EPA Meterological Wind Tunnel - Its Design, Construction

and Operating Characteristics," EPA-600/4-79-051, Environmental Sciences

Research Laboratory, Research Triangle Park, NC, September, 1979.

30. Greenway, M. E. and Wood, C. J. , "The Oxford University 4 m x 2 m Industrial

Aerodynamics Wind Tunnel," Journal of Industrial Aerodynamics , Vol. 4, No. 1,

January, 1979, pp. 43-70.

31. Morel, T. "Comprehensive Design of Axisymmertic Wind Tunnel Contractions,”

Transactions of the ASME, Journal of Fluids Engineering , Vol. 97, No. 1,

June, 1975, pp. 225-233.

32. Morel, T., "Design of Two-Dimensional Wind Tunnel Contractions," Transactions

of the ASME, Journel of Fluids Engineering, Vol. 99, No. 2, June, 1977,

pp. 371-378.

33. Streeter, V. L. , Fluid Mechanics, 2

n<* Ed., McGraw-Hill Book Co., Inc.,

New York, NY., 1958.

34. ASHRAE Handbook - 1981 Fundamentals , American Society of Heating,

Refrigerating and Air-Conditioning Engineers, Inc., Atlanta, GA., 1982.

35. Dougherty, R. L. and Ingersoll, A. C., Fluid Mechanics , McGraw-Hill

Book Co., Inc, New York, NY, 1954, pp. 191.

36. Engineering Sciences Data Unit, "Performance of Conical Diffusers in

Incompressible Flow," Data Item 73024, ESDU, London, March, 1980.

37. Engineering Sciences Data Unit, "Pressure Drop in Ducts Across Round-Wire

Gauzes Normal to the Flow," Data Item 72009, ESDU, London, June, 1972.

38. Engineering Sciences Data Unit, "A Guide to Fan Selection and Performance,"

Data Item 79037, ESDU, London, December, 1979.

34
Appendix I - NOTATION

A = Area; dimensionless coefficient

Aq = Area of test section

Cp = Pressure coefficient

Cpi = Wall pressure coefficient at inlet

Cp r = Static pressure recovery ratio

Cp r = Theoretical maximum pressure recovery ratio

Cr = Contraction ratio

D = Diameter

d = Characteristic dimension

E 0 = Efficiency of diffuser/screen combination

t E

q = Diffuser efficiency

E.R. = Energy ratio

F r = Froude number

g = Gravitational constant

H = Height of section

h = Height of model; contraction offset

I = Intensity of turbulence

K = Head-loss coefficient

Kt = Total-pressure loss coefficient

L = Length

n = Number of screens

nQ = Modal frequency

P = Static pressure

P t = Total pressure

q = Dynamic pressure

35
Rg = Reynolds number

U = Velocity

Ufo = Velocity at height of model

U-l = Velocity at inlet

UQ = Mean velocity in test section

Ur = Reduced velocity

Uz = Velocity at height z

Uco = Freestream velocity

U = Local mean velocity

u = Departure from mean velocity in longitudinal direction

u,v,w = x,y and z components of velocity fluctuations

V = Wall velocity

X = Distance in direction of flow

Y = Ordinate of contraction wall profile

Z = Height above test section floor

a = Kinetic energy coefficient

3 = Potential energy coefficient

6 = Boundary layer thickness

5* = Boundary layer displacement thickness

v = Kinematic viscosity

p = Mass density

36
Table 1 Contraction Geometry and Pressure Gradient

X/L Y/h Area

(ft) 2 c

p

dCp

dX

(ft)

0.000 1.000 192.000 0.902 0.000

0.050 0.986 189.652 0.900 0.007

0.100 0.948 183.312 0.893 0.013

0.150 0.891 174.033 0.881 0.020

0.200 0.819 162.802 0.864 0.028

0.250 0.738 150.497 0.841 0.038

0.300 0.652 137.853 0.811 0.050

0.350 0.563 125.456 0.771 0.063

0.400 0.475 113.749 0.722 0.079

0.450 0.391 103.037 0.661 0.095

0.500 0.313 93.516 0.588 0.112

0.550 0.242 85.282 0.505 0.126

0.600 0.179 78.359 0.414 0.134

0.650 0.127 72.718 0.319 0.134

0.700 0.084 68.287 0.228 0.125

0.750 0.051 64.968 0.147 0.105

0.800 0.027 62.638 0.083 0.079

0.850 0.012 61.155 0.037 0.050

0.900 0.004 60.356 0.012 0.024

0.950 0.001 60.046 0.002 0.007

1.000 0.000 60.000 0.000 0.000

1 ft = 0.3048 m

37
a) co •H

T—

o 0

•H J4 o

4-1 cO 3 o t-. p u rH

0-i •H o

40 Cu rH 3

rH

a Vj

cO <U 3 -a 3 TO

0) 5 J-i o -3 o O

•H o 3

Cfl 4-1 40

cO cO 3 4-1 M s 3 <f

43U

O

m roo in <r so 3 S co i£) eg

ai v-/ • • •

r-J cn s 44 o .3 o o O'

•H 00 /-s CO o eg U •H 0 rH o 00 o 4-) a) ’— • e • cO <y 03 eg eg rH

J44-) 3oou3S

•rH CD

a>

a)

j4

4-1o

eg

aj rH

40

CO H

40

O o 00

U esvO o <r

30 0 i£> o o •H • • • 13 co CO

44

O 00 00

40 M3 <r

m

00 ^ O'! 00 M3

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rc CO CO

4=O

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•H \-y • 3 in

00

eg

00m o

cn

hh i—I QJ

CO

oo o

CO •H 3 44

O o O -H rH «H rH H CO 44 • • • • CO J4 CO ON 00 CM 3 44 OS r^. • • <u c • CM CO S o CM

•rH o

TO co cu TO 3 3 3 3

•H T3 3 •H 3 3 3 H H 3

!—

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•h as •H *H

m a; oo eg CD 12 rH CD rH O V4 o 0

<u i—i • ai cd to a, 3 O 44 > o V4 o 00 3 j-i a> •H *H 3 J4 3 O oi 3 0 o H <u w 3 3 3 ' CO

44 3 3 • TO <U rH TO to O rH o 3 S <u J4 TO ea ai •H 3 o o 3 II 3 < 3 44 J4 cn e

PL, 3 X <U PQ 3 44 W H o <2 3 H 44

l““l rH CM CO 38
Table 3. Calculation of Losses and Energy Ratio. Closed-Circuit Tunnel

Tunnel Component

Ki Ai

(sq.ft)

Koi Percent of Total Loss

Ref. Dynamic Pressure

(psf) Loss(psf)

Total Pressure

(psf)

Static Pressure

(psf)

7.296 5.264 -2.032

Test Section 0.096 60 0.096 5.08 (33) 0.700

First Corner 7.104 4.564 0.460

Vanes 0.100 80 0.056 2.96 (13) 0.409

Duct 0.013 80 0.007 0.37 (33) 0.051

4.104 4.104 0 Transition and Silencer 0.500 64 0.439 23.24 (34) 3.203

6.413 0.901 -5.512

Second Corner Vanes 0.050 64 0.044 2.33 (13) 0.321

Duct 0.019 64 0.017 0.90 (33) 0.124

6.413 0.456 -5.957

Square-to-Round Transition 0.040 50.3 0.057 3.02 (35) 0.416

10.396 0.040 -10.356

Fan -13.782

10.396 13.822 3.426

Circular Duct 0.083 50.3 0.118 6.25 (33) 0.861

10.396 12.961 2.565

Round-to-Square Transition 0.130 50.3 0.185 9.79 (36) 1.350

6.413 11.611 5.198

First Wide-Angle Diffuser 0.180 64 0.158 8.36 (15,18) 1.153

3.103 10.458 7.355

Third Corner Vanes 0.050 96 0.020 1.06 (13) 0.146

Duct 0.016 96 0.006 0.32 (33) 0.044

3.103 10.268 7.16 5 Fourth Corner Vanes 0.050 96 0.020 1.06 (13) 0.146

Duct (including entry & exit 0.019 96 0.007 0.37 (33) 0.051

3.103 10.071 6.968

Second Wide-Angle Diffuser

Screen No. 1 0.700 108.7 0.213 11.28 (37) 1.554

2.223 8.517 6.294

Screen No. 2 0.700 148 0.115 6.09 (37) 0.839

1.199 7.678 6.479

Screen No. 3 0.700 192 0.068 3.60 (37) 0.496

0.713 7.182 6.469

Settling Chamber

Honeycomb/Screen 1.200 192 0.117 6.19 (23,24,37) 0.854

0.713 6.328 5.615 Duct 0.006 19 2 0.006 0.32 (33) 0.044

0.713 6.284 5.571 Contraction 0.040 60 0.040 2.12 (35) 0.292

7.296 5.992 - 1.304

Bluff Body in Test Section 0.100 60 0.100 5.29 (34) 0.730

1=1.889 £=100.00 7.296 5.262 - 2.034 |

Notes: 1 1 ft = 0.3048 m E.R. = 1 /JKq* = 0.53

2. Pressures are based on UQ

= 80 fps

3. 1 psf = 47.88 Pa

39
Table 4. Calculation of bosses and Energy Ratio. Open-Circuit Tunnel

Ki Ai Koi Percent of Dynamic Total Static

Tunnel Component (sq.ft) Total Loss Ref. Pressure Loss Pressure Pressure

(psf) (psf) (psf) (psf)

0 0 0 Fan — 2.610

-18.751

18.751 16.141

Plenum 100.3 — 2.369

0 18.751 16.382

First Diffusing Screen 2.800 105.3 0.909 35.38 (37)

2.369

6.631

12.120 9.751

Wide-Angle Diffuser 0.130 10 5.3 0.042 1.63 (36)

0.713

0.306

11.814 11.101

Second Diffusing Screen 2.800 19 2 0.273 10.62 (37)

0.713

1.992

9.822 9.109

Settling Chamber

Honeycomb/Screen 1.200 192 0.117 4.55 (23,24,37)

0.713

0.854

8.968 8.255

Duct 0.006 192 0.006 0.23 (33)

0.713

0.044

8.924 8.211

Contraction 0.040 60 0.040 1.56 (35)

7.296

0.292

8.632 1.336 Test Section

Bluff Body 0.100 60 0.100 3.89 (34)

7.296

0.730

7.902 0.606

Duct 0.083 60 0.083 3.23 (33)

7.296

0.606

7.296 0 Discharge 1.000 60 1.000 38.91 7.296

1=2.570 I = ioo 00 0 0 0

E.R. = 1/IKqI = 0.39

1. 1 ft = 0.3048 m

2. Pressures are based on U0 =

3. 1 psf - 47.88 Pa

Notes

80 fps
Square-to-Round

Transition

13

Second

Wide-Angle

Diffuser

Drive

Section

14

Settling

Chamber

Return

Duct

15

Contraction

o-.ez

iH CM CO -3- m vO r»

9Uj| ||BAA

co

T-l

>

<u rHw

cfl Co

•Ho

<u

C/5 -oc

tfl c

0

1 TJ

<U

00ou

a>u

3

oo

bu

41

SECTIONAL

ELEVATION
Figure 2. Turning Vane Profile. Closed-Circuit Tunnel

42
Chord

=

2.828

ft

£

width

=

3.206

Radius

=

1.564

Pitch

=

0.500

43

Figure

3.

Typical

Vane

Cascade.

Closed-Circuit

Tunnel
Figure 4. First Wide-Angle Diffuser. Closed-Circuit Tunnel

44
45

Figure

5.

Second

Wide-Angle

Diffuser.

Closed-Circuit

Tunnel
Figure 6. Settling Chamber
3

47

Figure

7.

Contraction

Profile.

Definition

Sketch
48

Figure

8.

Pressure

Coefficient

and

Gradient

for

Three

Similar

Contractions
l8|jnO UOliOBJlUOO

japno jaqweijo Buipias

lapno jasnjjip puooas

lapno jaujoo qi-inoj

ja|U| jaujoo qt-inoj

iapno jaujoo pj|m

lapno jasnjpp jsjij

japno punoj-oj-ajenbs

japno ionp JB|noj[Q

jaiino ubj

) d|u; ubj

lapno jaujoo puooas

jai;no jaouans

iapno jaujoo jsjij

uoijoas jsaj puj

o

T- | (jsd) 3anSS3dd (jsd) 3dnSS3dd

<uc

c

3H

3u

u

•Ho

4>

<u

COo

c_>

o

>-<

0-

i

a>u

3

CO09

1)

1-

i Cl.

O'

11

1-. 3

00 H

U,

49

Note:

Test

section

not

shown
TUNNEL

SPEED

(fps)

(J01BM jo soipui) SSCH QV 3 H

50

Figure

10.

Tunnel

and

Fan

Characteristics.

Closed-Circuit

Tunnel
Figure 11. Power Demand Curve. Closed-Circuit Tunnel

51
Wall

line

1 Centrifugal Fan

2 Plenum

3 Wide-Angle Diffuser

4 Settling Chamber

5 Contraction

6 Test Section

Figure 12. Open-Circuit Tunnel. Plan and Elevation

52

Wall

line
4.400’

I

I

5.700'

Diffusing screen Diffusing screen

Figure 13. Wide-Angle Diffuser. Open-Circuit Tunnel

53

6

.

000

'

8.000
PRESSURE

(psf )

PRESSURE

(psf)

20

15

10

5

0

c

<D 0 £ ~ O

3 (0 o

C W)

CO .!= Ll IJL

I n

/

/ r /

/ 7

/

0

3

o

l_0

w

3

0

JQ

E

03 JC 0

CD 2

0

CO

±

_0

-t-* 3

o

c

o

o

0

*-> c

o

o

TOTAL PRESSURE

1L.

_L JL ± U

20 40 60 80

DISTANCE FROM FAN INLET (ft)

100

Figure 14. Pressure Profiles. Open-Circuit Tunnel. (UQ = 80 fps)

54

Tunnel

outlet
TUNNEL

SPEED

(fps) ooo

o'o

COooo

o'o

CMoo

o^

o'oo 2

LLo

>-o<

CL<o

55

Figure

15.

Tunnel

and

Fan

Characteristics.

Open-Circuit

Tunnel
Figure 16. Power Demand Curve. Open-Circuit Tunnel

56
U.S. DEPT. OF COMM. 1. PUBLICATION OR 2. Performing Organ. Report No. 3. Publ cation Date BIBLIOGRAPHIC DATA

SHEET (See instructions)

REPORT NO.

NBSIR 85-3168 May 1985

4. TITLE AND SUBTITLE

Performance Requirements and Preliminary Design of a Boundary Layer Wind Tunnel

Facility

5. AUTHOR(S)

Richard D. Marshall

6. PERFORMING ORGANIZATION (If joint or other than NBS, see instructions)

NATIONAL BUREAU OF STANDARDS

7. Contract/Grant No.

DEPARTMENT OF COMMERCE 8. Type of Report & Period Covered

Gaithersburg, Final MD 20899

9. SPONSORING ORGANIZATION NAME AND COMPLETE ADDRESS (Street. City. State, ZIP)

same as item 6 above

10.

SUPPLEMENTARY NOTES

[ ] Document describes a computer program; SF-185, FIPS Software Summary, Is attached.

11.

ABSTRACT (A 200-word or less factual summary of most significant information. If document includes a significant

bi bl iography or literature survey, mention it here) This report describes performance characteristics and design details of a boundary

layer wind tunnel for supporting research activities within the Center for Building

Technology. Two preliminary designs, the first consisting of a conventional closedcircuit scheme in an over/under configuration and the second consisting of an opencircuit scheme with a "pusher" or "blow-down" configuration, are addressed. Both

tunnels incorporate a three-dimensional contraction with a 3.2:1 contraction ratio

and a test section having a width of 10 ft (3.048 m) and a height of 6 ft (1.829 m) . The test section length allows for the natural development of rough-wall boundary

layers of sufficient depth for 1:500 scale simulations of the atmospheric boundary

layer. To ensure stable operation over the design speed range of 1 to 75 fps

(0.30 to 22.86 m/s), both variable pitch and variable speed fan control will be

required.

12.

KEY WORDS (Six to twelve entries; alphabetical order; capitalize only proper names; and separate key word-. : , Aerodynamics; boundary layers; buildings; modeling; wind engineering; wind tunm-!' .

13. AVAILABILITY 14. NO. OF

PRINTED Sumimited

1 | For Official Distribution. Do Not Release to NTIS

l -1 Order From Superintendent of Documents, U.S. Government Printing Office, Washington, D.C. 20402.

67

15. Price

PAGE

-

Order From National Technical Information Service (NTIS), Springfield, VA. 22161

UU * f