IMPROVING THE PERFORMANCE OF AN AC UNIT USING CFD

Size: px

Start display at page:

Download "IMPROVING THE PERFORMANCE OF AN AC UNIT USING CFD"

Steven Dorsey
5 years ago
Views:

1 3 rd HVAC Contracting Conference Evolution of HVAC Industry International Hotel Doha The City November 25-26, 2018 IMPROVING THE PERFORMANCE OF AN AC UNIT USING CFD F. Moualled Mechanical Engineering Department American University of Beirut 1

2 Objectives The objectives of this presentation are: To present a numerical model developed for predicting the performance of AC equipment. To use the model for analyzing the performance of a 10-ton Rooftop Unit. To chec the practicability of using CFD to complement experimental wor and to help optimizing the performance of AC equipment. 2

3 Schematic of the AC Unit 3

4 Unit Description and Refrigeration Cycle The nominal refrigeration capacity of the unit is 10 tons of which 7 tons are sensible. The volume flow rate is 4000 cfm with a fan speed of 770 rpm. The volume flow rate in the condenser is 7200 cfm with a fan speed of 1090 rpm. 4

5 Modeling of the Unit The unit is divided into two compartments with no air communication between them: The evaporator section: The surface temperature of the evaporator being lower than the air dew point temperature causes condensation of its moisture. Due to condensation, the air flow side in the evaporator is modeled as a two-phase flow. The condenser section :The air flow side in the condenser is modeled as a single phase flow. Communication between these two compartments occurs at the refrigerant level. 5

6 Governing Equations The conservation equations governing the flow, heat, and mass transfer fields in the unit are: Conservation of mass Conservation of momentum Conservation of energy Conservation of vapor mass fraction Conservation of liquid volume fraction Turbulence model (Realizable -e) 6

7 7 Governing Equations 0 m m m t u dr dr ij m m m m p t,, u u F g u u u h eff S T p E E t u v v v v S Y t Y J u v t t m v v Y Sc D, J S t dr m, u u b t S G G t e u e e e e e e e e e e S G C C C S C t b t u

8 Condensation Model cold wall m cond C v D v, m D t Y Distance v, cell Y sat, wall centroid cell, centroid wall cell centroid Q lat mcond Lvap c p, vaptcell c p, liq T wall wall centroid Y v =Y sat(t,p). mcond Yv 8

9 CFD Modeling of The Evaporator 9

10 10

11 11

12 Boudary Conditions Evaporator No slip conditions at the wall with a surface temperature of 282 K. Centrifugal Fan Modeled as a moving frame of reference with rotational speed of 770 rpm. Fresh and Return air inlets Modeled as velocity inlets with a uniform normal velocity of value 3.99 m/s and a temperature of 300 K. Outlet Modeled as a pressure outlet with temperature of K and gauge pressure of 37Pascal. Motor wall Modeled as a source of heat flux where it is defined as wall with flux of 888 W/m 2. Walls No slip condition is assumed where the velocity at the walls is taen to be zero. 12

13 Results Velocity Vectors 13 Isotherms

14 Results Isobars Water mass fraction contours 14

15 Results Pathlines of particles released along a straight line from the inlet section

16 Table 1 Comparison of measured and computed cooling capacities (W) and condensation rate (l/hr ) for a volume flow rate of 1.89 m 3 /s (4,000 cfm) and indoor and outdoor dry bulb temperatures of K (80 F) and K (95 F), respectively. Wet bulb temperature K (62 F) (T coil = K) K (67 F) (T coil = K) K (72 F) (T coil = K) Type of Prediction Exp. Num. %Diff Exp. Num. %Diff Exp. Num. %Diff Total Cooling Capacity 33,377 32, ,610 36, ,185 40, Sensible Cooling 31,817 30, ,373 26, ,913 21, Capacity Latent Cooling Capacity 1,561 1, ,237 10, ,273 19, Computed condensation rate Table 2 Comparison of measured and computed cooling capacities (W) and condensation rate (l/hr ) for a wet bulb temperature of K (67 F) and indoor and outdoor dry bulb temperatures of K (80 F) and K (95 F), respectively. m 3 /s (cfm) 1.51 (3,200, T coil = K) 1.89 (4,000, T coil = K) 2.27 (4,800, T coil = K) Type of Prediction Exp. Num. %Diff Exp. Num. %Diff Exp. Num. %Diff Total Cooling Capacity 35,014 35, ,610 36, ,778 38, Sensible Cooling 23,715 25, ,373 26, ,780 28, Capacity Latent Cooling Capacity 11,300 10, ,237 10, ,998 10, Computed condensation rate

17 Modifying Evaporator Design x d W y 2y+d x/2 H L x/2 2x+d 17

18 Modifying Evaporator Design Coil Original Design 1 Design 2 Design 3 Design 4 Design 5 Design 6 H (mm) L (mm) W (mm) Inclination from vertical Reduction in evaporator size Vertical distance between pipes, x (mm) Horizontal distance between pipes, y (mm) Face coil average velocity % 0 15% 10% Number of pipe rows Total number of pipes Number of fins per inch Total number of fins Pipe diameter (mm) Total coil area (m 2 )

19 Performance of Newly Suggested Designs Type of Prediction Original 1 st %Diff 2 nd %Diff 3 rd %Diff 4 th %Diff 5 th %Diff 6 th %Diff Total Cooling Capacity 36,347 design design design design design design 37, , , , , , Sensible Cooling Capacity Latent Cooling Capacity 26,341 10,006 26, , , , , , , , , , , , Computed condensation rate 19

20 CFD Modeling of The Condenser 20

21 Refrigeration Cycle of the Rooftop AC Unit T (K) 380 p=21.95 bar h=314 J/g p=21.59 bar h=103 J/g p=5.08 bar h=55 J/g 1 p=5.09 bar h=253 J/g S 21

22 Results Velocity Vectors 22 Pathlines

23 Results Isotherms 23 Isobars

24 Results The calculated total heat transfer rate of the condenser coil is found to be 44,119.5 W while the calculated value based on the experimental data provided is 44,120.1 W with the percent difference between the two values being %. 24

25 CONCLUSIONS Based on the computations performed for the suggested designs the following conclusions were drawn: (i) Decreasing the inclination angle increases the cooling capacity at the expense of increasing the air velocity; (ii) Bacward inclination of the cooling coil reduces its cooling capacity; (iii) Increasing the number of fins per unit length increases the cooling capacity; (iv) The best improvement in performance is obtained with decreasing the inclination angle and increasing the number of fins per unit length; (v) The rate of increase in cooling capacity is affected more by the number of fins than by the inclination angle. Based on the results obtained in this study it can be safely stated that CFD will be a viable tool in the design of AC equipment. Finally, it should be mentioned that the above findings remain an open question for the industry to confirm. 25

26 THANK YOU FOR YOUR ATTENTION Questions? 26

International Journal of Heat and Mass Transfer

International Journal of Heat and Mass Transfer 54 (2011) 549 563 Contents lists available at ScienceDirect International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt