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FLUID FLOW REGIMES IN AIR CONDITIONED HALL WITH HEAT SOURCE EQUIPMENT

ASSER EL-ERIAN and ESSAM E. KHALIL

Faculty of Engineering, Cairo University Cairo –EGYPT

ABSTRACT The present paper deals with experimental and numerical investigations of the influence of ventilation and air Conditioning supply and extract openings locations on room flow and thermal patterns in buildings. This work focuses on mechanical ventilation in industrial building. The effectiveness of a ventilation system is mainly determined by the removal of internally produced contaminants from the room (if any) and the supply of fresh air of acceptable quality in the room to replace contaminated air and reduce thermal load in inhabited zone. This is targeted to attain desired temperature regulation for human thermal comfort and machine safe operating thermal conditions. The present experimental facilities include a main room which is part of the pumping station complex having principal dimensions of 9 m long x 8m wide and 7m high. It includes mainly an electric motor driven screw pump; mounted on a concrete skid . The air supply comprises two short-ducted axial fans 0.483m diameter. The Extracted air side consists of a single roof extraction fan having the diameter of 0.76m with a specified fan duty of 3.82 m3/s. The main heat source has the overall dimensions of 1.5 m x 2.0m wide and 2.m high. Platinum Resistance thermometers were employed to measure local temperatures in the room and vicinity of heat source. On the other hand air velocity measurements were as low as 0.01 m/s and the turbulence intensity as high as 10%. The measuring probe, however, is omni-directional, the inaccuracy was in the order of 10 - 20% at 0.05 - 0.5 m/s and was regularly calibrated. The heat sources in the room create buoyant plumes that transport the air from the lower part (direct above the heat source) into the higher zones through entrainment. The velocity distributions have been measured at same vertical planes above the heat source similar to those of the temperature measurements. A numerical study was carried out to define the optimum airside design of the HVAC systems that provides the optimum ventilation and energy utilization. The present model is packaged as a Computational Fluid Dynamics (CFD) program and is named under the title 3DHVAC; a non-uniform grid of size of minimum of 80 X 60 X 44 was selected. Grid nodes were densely located in the vicinity of the heat source. The primary objective of the present work is to assess the airflow characteristics and Heat transfer in large air-conditioned configurations with large heat source; reasonably good agreement between measured and predicted air velocities and temperatures were reported. The paper ends with a brief discussion and conclusion.

44th AIAA Aerospace Sciences Meeting and Exhibit9 - 12 January 2006, Reno, Nevada

AIAA 2006-1509

Copyright © 2006 by the American Institute of Aeronautics and Astronautics, Inc. All rights reserved.

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INTRODUCTION To design an optimum HVAC airside system that provides comfort and air quality in the air-conditioned spaces with efficient energy consumption is a great challenge. Air conditioning can be identified as the conditioning of the air to maintain specific conditions of temperature, humidity, and dust level inside an enclosed space. The conditions to be maintained are dictated by the function of the space, type of users and the required users comfort. So, the air conditioning embraces more than cooling or heating. The comfort air conditioning is defined as “the process of treating air to control simultaneously its temperature, humidity, cleanliness, and distribution to meet the comfort requirements of the occupants of the conditioned space”1 .This definition includes the entire heat exchange operation, regulation of velocity, thermal radiation and quality of air, as well as the removal of foreign particles and vapors. 2 A successful HVAC design is that energy efficient design in addition to all previous factors For the present work, following other earlier work 3-5, a numerical study is carried out to define the optimum airside design of the HVAC systems, which provides the optimum comfort and healthy conditions with optimum energy utilization. The present model is packaged as a Computational Fluid Dynamics (CFD) program and is named under the title 3DHVAC 3, 4. The present paper introduces a description of the present 3DHVAC program and its validation with steady state results from the literatures. Basically, various airside designs are considered here including floor supply, different obstacle and alternative positioning to introduce the capability of each design to provide the optimum air flow characteristics. The primary objective of the present work is to assess the airflow characteristics and energy consumption in the different air-conditioned configurations in view of basic known flow characteristics. The paper ends with a brief discussion and conclusion.

PROBLEM FORMULATION The proper efficient airflow distribution is required in all applications in their both categories; residential and industrial. The airflow distribution in its final steady pattern is a result of different interactions as shown in Figure 1, such as the airside design, equipment distribution, thermal effects, occupancy movements, etc. The airside design and internal obstacles are the focus of the present work. The supply outlets and extraction intakes play an important role in the main flow pattern and the creation of main recirculation zones. The internal obstacles can affect the airflow pattern by different ways, such as, by increasing the recirculation zones or by deflecting the main airflow pattern. METHOD DESCRIPTION MODEL EQUATIONS The program solves the differential equations governing the transport of mass, three momentum components, energy, relative humidity, and the air age (residence time) in 3D configurations under steady conditions. The different governing partial differential equations are typically expressed in a general form as:

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ΦΦΦΦ +⎟⎟⎠

⎞⎜⎜⎝

⎛∂Φ∂

Γ∂∂

+⎟⎟⎠

⎞⎜⎜⎝

⎛∂Φ∂

Γ∂∂

+⎟⎟⎠

⎞⎜⎜⎝

⎛∂Φ∂

Γ∂∂

=Φ∂∂

+Φ∂∂

+Φ∂∂ S

zzyyxxW

zV

yU

x effeffeff ,,,ρρρ ………(1)

ρ = Air density, kg/m3 Φ = Dependent variable. SΦ = Source term of Φ. U, V, W = Velocity vectors. ΓΦ,eff = Effective diffusion coefficient. The effective diffusion coefficients and source terms for the various differential equations are listed in the following table.

Table 1: Terms of Partial Differential Equations Partial Differential Equations

Φ ΓΦ,eff SΦ Continuity 1 0 0 X-momentum U µeff -∂P/∂x+ρg + SU Y-momentum V µeff -∂P/∂y+ρg + SV Z-momentum W µeff -∂P/∂z+ρg(1+β∆t) +SW H-equation H µeff/σH SH RH-Equation RH µeff/σRH SRH τ-age equation τ µeff/στ ρ k-equation k µeff/σk G - ρ ε ε-equation ε µeff/σε C1 ε G/k – C2 ρ ε2/k µeff = µlam + µ t µ t = ρ Cµ k2 / ε G = µt [2{(∂U/∂x)2 +(∂V/∂y)2 +(∂W/∂z)2}+(∂U/∂y +∂V/∂x)2 +(∂V/∂z + ∂W/∂y)2 +(∂U/∂z + ∂W/∂x)2] SU = ∂/∂x(µeff ∂Φ/∂x)+∂/∂y(µeff ∂Φ/∂x)+∂/∂z(µeff ∂Φ/∂x) SV = ∂/∂x(µeff ∂Φ/∂y)+∂/∂y(µeff ∂Φ/∂y)+∂/∂z(µeff ∂Φ/∂y) SW = ∂/∂x(µeff ∂Φ/∂z)+∂/∂y(µeff ∂Φ/∂z)+∂/∂z(µeff ∂Φ/∂z) Turbulence model constants C1 = 1.44, C2 = 1.92, Cµ = 0.09 σH = 0.9, σRH = 0.9, στ = 0.9, σk = 0.9, σε = 1.225

Figure 1: Graphical representation of room zoning, according to the prevailing flow regimes COMPUTATIONAL GRID GENERATION The present program utilizes the modified hyperbolic formula 6 for grid node generation. The presently used formula creates a non-uniform orthogonal grid with dense gird nodes near the walls and internal obstacles. The present program is also designed to simulate

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the airflow domain using 384,000 cells to obtain grid independent predictions with efficient cost. This number of cells gives the tolerance to enhance the flow representation in the air outlets and intakes. NUMERICAL SCHEME The present program can utilize either the power-law scheme that was proposed and recommended 7 or alternatively the simple first-order upwind difference scheme8 . Although the power-law scheme is only first-order accurate on the basis of truncation error, the power-law difference scheme is a conservative formulation and does not have the problem of numerical oscillations. The drawback of the scheme is the inherent numerical diffusion. In the recirculating zones, the effective diffusion will be replaced by numerical cross-flow diffusion in regions of the solution domain where the flow is not aligned with the grid-lines 8, 9.Upwind Difference Scheme is used here. INITIAL AND BOUNDARY CONDITIONS The initial conditions of the solving domain are taken to represent the actual value of each variable Φ . The boundary conditions represent the wall, inlet, and outlet conditions. The detailed treatment of boundary conditions can be found in the literature that describes the 3DHVAC program 4. Near the wall, the log-law of the wall function is applied to correct and find the values of turbulence and dissipation. The velocity of the airflow tends to be zero at the wall surface. The effect of buoyancy forces was included. CONVERGENCE AND STABILITY The simultaneous and non-linear characteristics of the finite difference equations necessitate that special measures are employed to procure numerical stability (convergence): these include under relaxation of the solution of the momentum and turbulence equations by under relaxation factors which relate the old and the new values of Φ as follows: ( ) oldnew 1 Φγ−+Φγ=Φ …………………………………….…………….…… (2) Where: γ is the under-relaxation factor. It was varied between 0.2 and 0.3 for the three velocity components as the number of iteration increases. For the turbulence quantities, γ was taken between 0.2 and 0.4 and for other variables between 0.2 and 0.6. The required iterations for convergence are based on the nature of the problem and the numerical conditions (grid nodes, under-relaxation factor, initial guess, etc.). So the time (on the computer processor) required to obtain the results is based on many factors. The computational number of iterative steps is selected according space cell (spatial difference) to yield converged solutions9.The room boundaries are represented as constant temperature boundaries. A constant heat flux is prescribed for the electric motor coupled to a gear box. Identical boundary conditions are used for the opposing walls. The standard k-ε model has been used to model the turbulence and standard wall functions have been applied to represent the near wall flow field.

EXPERIMENTAL INVESTIGATION The main room is part of the pumping station complex having principal dimensions of 9.0m long x 8.0m wide and 7.0m high. It includes mainly an electric motor coupled to a gear box (as one unit) which drive the screw pump all mounted on a concrete skid as shown in Figure 5.1. The room ventilation system comprises of two supply opening and single exhaust. The inlet side comprises of two short duct supply axial fans having the diameter of 0.483m. Each fan is mounted horizontally, running with speed of 1450 RPM and having a specified fan duty of 1.91 m3/s. The Exhaust side consists of a single roof

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extraction fan having the diameter of 0.76m with a specified fan duty of 3.82 m3/s.The main heat source in the room is the electric motor with integrated gear box. The electric motor, 130HP with the integrated gear box and the concrete skid as whole unit has the dimensions of length 1.5m x 2.m wide and 2.m high. MEASUREMENTS The measurements were carried out in main room of the pumping station complex concern airflow velocity, air temperature. Platinum Resistance thermometers were employed to measure local temperatures in the room and vicinity of heat source. On the other hand air velocity measurements were as low as 0.01 m/s and the turbulence intensity as high as 10%. The measuring probe, however, is omni-directional, the inaccuracy was in the order of 10 - 20% at 0.05 - 0.5 m/s and was regularly calibrated. The heat sources in the room create buoyant plumes that transport the air from the lower part (direct above the heat source) into the higher zones through entrainment. The velocity distributions have been measured at same vertical planes above the heat source similar to those of the temperature measurements. The pumping station room was simplified to be an empty room with two air supplies and a single exhaust containing only the gear box, electric motor the concrete skid which all were consider as a single solid block emitting heat as in figure 2 . As measurement first principle, the desired area of study (source zone) was chosen in particular the expected plume zone above the heat source (eclectic motor with integrated gears) which helps to assess in the selection of optimum measurement locations that will aid in well description of the desired phenomena. The heat sources in the room create buoyant plumes that transport the air from the lower part (direct above the heat source) into the higher part through entrainment. From the full-scale measurements information is available on the plume above the thermal source where temperature values were gathered form vertical plane above thermal source to bestow vertical temperature profile. For both units the velocity and the velocity distribution have been measured at same planes vertical above the heat source as in the temperature case.

Figure 2 Pump station room (drawing not to scale)

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VALIDATION PROCEDURE This section presents a sample of the validation results of the developed program when used on selected applications, previously investigated and reported by others; the observations that were noted during the validation process are discussed10. The validation cases were previously reported by the authors in reference 10 .These included the cases of Nielson 11 where extensive velocity measurements in a rectangular parallelepiped scaled model of a room (H = 89.3 mm) in which the isothermal airflow is expected to be almost two-dimensional was carried out 11 as shown in Figure 3. The length of the model (L = 3H). The condition parameters are based on inlet Reynolds number (Re = 5000) 11. Air enters through an opening at the top left wall and is discharged from the right wall at near floor level. In the present work, the same configuration was simulated using 3DHVAC and comparisons are presented here in the Figure 4. The comparisons indicate good agreement especially in the jet downstream with discrepancies of the order of 5 % – 7.5 %. The comparisons also show some discrepancies in the prediction at the recirculation zones, especially near the floor. Other comparisons with the experiments of Manzoni and Guitton 12 were shown to be qualitative as indicated in reference 10.

Figure 3: Predicted Airflow Pattern

Comparison at X = H Comparison at X = 2H

Figure 4: Comparison with Experiments [11]

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EXPERIMENTAL RESULTS AND SIMULATION MODELING

The flow regimes resulted from the air supply / return diffusers incorporated in the present case shown in Figure 2 was measured and the typical velocity characteristics were experimentally obtained and shown in Figure 5 together with the present Computations termed CFD13 . Good agreement is shown and the numerical simulation predicts the correct trend with discrepancies of the order of 10%. On the other hand ,the thermal pattern is shown in Figure 6 in terms of measured and predicted air temperatures. Figures 7 and 8 demonstrate the capabilities of the present modeling technique to readily yield more comprehensive details of the flow and thermal patterns.

2.22.32.42.52.62.72.82.9

33.13.2

0 0.2 0.4 0.6 0.8

Velocity (m/sec)

Hei

ght (

m)

CFDExp.

Figure 5: The velocity profile at vertical plan above heat source

2

2.2

2.4

2.6

2.8

3

3.2

3.4

3.6

3.8

23 24 25 26 27 28

Temperature ( C)

Hei

ght (

m)

Exp.CFD

Figure 6: The temperature profile at vertical plan above heat source

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DISCUSSION AND CONCLUSION

The present work reported some experimental investigation in a typical pumping station where heat is dissipated from the pump-motor assembly and is taken away by the air flow to cool down the room. One of the present goals is to validate 3DHVAC developed earlier by Kameel and Khalil14 for the present flow configurations and to examine how accurate the program 3DHVAC can provide an effective tool for displacement and mixing ventilation systems description. Mean Flow velocity was plotted against height in a plan over the heat source as these measured velocity values were compared and then examined; the numerical simulation yielded in underestimated vertical component of the velocity relative to measured values. Such discrepancy can be attributed to the difference between measured and predicted temperatures directly affecting the buoyancy term in the momentum equations. Nevertheless they both had the same profile and trend with good degree of accuracy. It can be seen that the Graphical representation in figure 1 that subdivided space into different zones, according to the prevailing flow regimes gives a good match with velocity field mapping. There are two zones above and below the inlet air supply that appears to comply with two recirculation zones also above and below the inlet air jet. Also the source zone that contains all of the inlet /outlet air arrangement plus the heat The vertical plan Y-Z at X=6.5 m was selected to present the temperature field in the calculated temperature contour mapping as in figure 8. The mixing ventilation temperature profile indicated that the room air flow could not completely mix in some parts of the ventilated space. Consequently a steeper short vertical velocity profile can be developed in ventilation mechanism as in figure 5. Figure 9 shows graphical flow features of the mixing ventilation flow pattern in the vertical plane of the room. A characteristic of a displacement ventilation flow pattern in a room is the relatively sharply located interface between the supply air layer, indicated the lower layer, and the upper layer .An interface is found at the height where the supply flow rate equals the convection flow rate in the plume(s). In the case of mixing ventilation as in both figures 8 and 9 the air is thoroughly mixed where the air layer immediately above the floor is cooled by the cold floor. Near the walls a downward flow transports air into the lower room level. The air near floor level is predominantly transported into higher room levels through entrainment into the warm air flowing upwards. With decreasing buoyancy near the walls, the air might circulate at several positions as it being entrained in warm upward flow.

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Figure 7: Calculated Velocity contour field, m/s at the centre plane of the room ,X and Z in meters.

Figure 8: Calculated Temperature contours(oC) field at the vertical plane of the room at X=6.5m , Y and Z in meters.

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Figure 9 :Graphical flow features of the displacement ventilation flow pattern.

REFERENCES [1] ASHRAE Handbook, Fundamentals 2001, ASHRAE, Atlanta, USA. [2] Stoecker, W. F., and Jones, J. W., 1985, Refrigeration and air conditioning, Second Edition, TATA McGraw-Hill Publishing Company LTD., 1985. [3] Khalil, E. E., 2000, Computer aided design for comfort in healthy air conditioned spaces, Proceedings of Healthy Buildings 2000, Finland, Vol. 2, Page 461. [4] Kameel, R., 2002, Computer aided design of flow regimes in air-conditioned operating theatres, Ph.D. Thesis work, Cairo University. [5] Kameel, R., 2003, Assessment of airflow characteristics in air-conditioned spaces using 3D time-dependent CFD model, IECEC 2003-5960, Portsmouth. [6] Kameel, R., and Khalil, E. E., 2002, Generation of the grid node distribution using modified hyperbolic equations, 40th Aerospace Sciences Meeting & Exhibit, Reno, Nevada, AIAA-2002-656, 12-15 January 2002. [7] Patankar, S. V. 1980, Numerical heat transfer and fluid flow, Hemisphere Publishing Corporation, WDC. [8] Sorensen, D. N., and Nielsen, P. V., 2003, Quality control of computational fluid dynamics in indoor environments, Indoor Air, Volume 12 No.1, page 2. [9] Leonard, B. P., and Drummond, J. E., 1995, Why you should not use ‘hybrid’, ‘power-law’ or related exponential schemes for connective modeling – there are much better alternatives, International Journal for Numerical Methods in Fluids, 20, 421-442, 1995.

[10] ElErian,A. and Khalil,E.E. Fluid Flow Regimes in Air Conditioned Rooms”,AIAA 2004-5593,Proc.2nd IECEC,RhodeIslans,USA,August 2004.

[11] Nielsen, P. V., Restivo, A. and Whitelaw, J. H., 1978, The velocity characteristics of ventilated rooms, J. Fluids Eng., 100, 291–298.

[12] Manzoni ,D. and Guitton , P. “Validation of displacement ventilation simplified modals”, Building Simulation 97,paper I-233,1997.

[13] ElErian,A., 2006 ,” Fluid Flow Regimes In Air Conditioned Hall With Heat Source Equipment”, MSc.Thesis, Cairo University,2006.

[14] Kameel, R., and Khalil, E. E., 2003, Energy efficiency, air quality, and comfort in air-conditioned spaces, DETC2003 / CIE – 48255, ASME 2003 Design Engineering Technical Conferences and Computers and Information in Engineering Conference, Chicago, Illinois USA, 2003.