Effect of Inlet on Efficiency and Flow Pattern in Centrifugal Fan using CFD Analysis and Experimental Validation

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1 Effect of Inlet on Efficiency and Flow Pattern in Centrifugal Fan using CFD Analysis and Experimental Validation MOJTABA GHOLAMIAN Department of Mechanical Engineering, JNTUH College of Engineering, Hyderabad, INDIA phone: , GURRAM KRISHNA MOHANA RAO * Department of Mechanical Engineering JNTUH, College of Engineering, Hyderabad, INDIA kmrgurram@jntuh.ac.in, ph: PANITAPU BHRAMARA Department of Mechanical Engineering JNTUH, College of Engineering, Hyderabad, INDIA bhramara74@yahoo.com *(Corresponding author) Abstract-Squirrel cage fans are special types of turbo machines commonly used in industries and buildings as a part of HVAC systems. These equipments with respect to other turbomachines use relatively lower energy. Since in industry and large buildings a group of them is used, consumed energy and also energy saving is considerable. Even though there are many important parameters that can affect the performance and efficiency (energy consuming), enough research has not been done on these parameters. Some experiments showed that there are some turbulent areas in volute and impeller and also in few cases that are reported in the literature, the relationship between the relevant parameters and turbulent areas or performance were defined. In this paper the effect of inlet nozzle and inlet diffuser on efficiency, performance and flow pattern is considered. Three inlet nozzle sizes with respect to impeller size were chosen (smaller, nearly same and bigger than inner impeller diameter) and numerical simulations were performed for both inlet shapes of nozzle shape and diffuser shape to find the effect of inlet shape on flow pattern, performance and efficiency. For the validation of numerical results, some experiments were done and all the performance parameters were compared that of numerical s. These results showed good matching between experimental and numerical results. Also these results show that inlet parameters, size and shape, have the grate effect on efficiency and flow pattern. At following the reason of efficiency differences, flow pattern and its mechanism are studied and presented. Key Words: Squirrel cage fan, Forward blades, Inlet Nozzle, Inlet Diffuser, Efficiency 1 Introduction Forward centrifugal fans due to their simplicity, ease of manufacturing and low costs are used extensively in industries and HVAC applications. This type of turbo machine has small chord line length, high width and blades that are joined together with shroud and hub. Figure1 shows a schematic sketch of the squirrel cage fan. Figure1: Schematic of squirrel cage fan ISBN:

2 Experiments and studies have shown that the flow in this equipment is three dimensional, complex and turbulent, especially in two regions: around one third of impeller width and near tongue in volute. This turbo machine has numerous geometrical parameters that can affect its performance. If a change in a parameter can improve the performance (especially efficiency), it means the flow pattern is more uniform and losses are reduced in impeller, particularly at the blade entry, between blades or at the exit of blades or volute or both of them. In centrifugal fans, flow enters from inlet almost uniformly and asymmetric [1] and goes into impeller. Measurements have shown that the flow cannot turn and change its direction from axial to radial at the first part of impeller width. Flow enters into the blade passage with high attack angle and it separates from the suction surfaces of blades and finally some parts of blades will be stalled. Flow at the rotor exit is three dimensional, highly fluctuating with jet and wakes. It is reported that in the area between inlet nozzle and up to one third of impeller width there is flow separation, high turbulent flow and reversed flow [2, 3]. Volute, impeller and inlet are basic elements in fans that have vital degree of importance in centrifugal fans. For impeller and volute some case studies (considerations and modifications) have been done to find out and modify the interaction between impeller and volute and flow pattern at exit of impeller and also near tongue [4, 5, 3, 6]. Yu-Taj Lee [5] used CRUNCH CFD code with K- Ԑ equation for turbulent modeling and Wall- Function procedure and showed that there is a well agreement between experimental and numerical method for static pressure, but not for power. Sofiane Khelladi [3] studied flow in impellerdiffuser interface by using FLUENT code with k-ω SST (Shear Stress Transport) model and showed that there is a well agreement between numerical and experimental results for pressure in impeller inlet and return channel outlet, but Kwang Yong Kim [7],to do shape optimization for centrifugal fans, and Rafael [8],considered a special case in which only the radial gap and clearance between inlet nozzle and impeller is studied, used numerical methods with k-ԑ model, SIMPLE and SIMPLE C algorithm and finally they showed that there is a well agreement between numerical and experimental results for performance in different flow rates. Here, for validation of steady state solutions and simulations, experimental cases were performed.inlet is an important element of fan that has some parameters and can have a great effect on efficiency and variation of losses especially between inlet nozzle exit and first section of impeller width. Inlet shape, nozzle shape and diffuser shape, is the basic parameter of inlet that can have a great role among the other inlet parameters. Most of researchers considered this parameter as a fixed parameter and other ones compared different fan performances of some literatures while there is not any information about the inlet shape. Under this condition, since other elements and parameters were not maintained unchanged, the corresponding results do not represent that the decreasing or increasing of losses (performance and efficiency) are only due to inlet nozzle of fan. To overcome the problems associated with the earlier researches, a case study is needed while all parameters are fixed and only the inlet diameter and inlet shape are varied. By this method the effect of inlet on performance may not be influenced by other parameters. In the present study with respect to the flow pattern from inlet to impeller [1, 4, 8*], three inlet sizes, towards achieving best performance and efficiency, with respect to impeller diameter, less than inner impeller diameter, near inner impeller diameter and finally greater than inner impeller diameter, at both shapes of inlet were chosen to perform numerical simulations. To validate the numerical results, some experiments were conducted with the same geometrical parameters as considered in the numerical simulations. Some of these dimensional and non-dimensional parameters are presented in Table1. ISBN:

3 Table 1- Parameters and Dimensions of the fan for simulation and experimentation at both shape of nozzle shape and diffuser shape D i 175mm, 190mm, 205mm α s D shroud 218mm ω 900rpm D 1 186mm α 30 o D 2 216mm C nozzle shape D 1 / D C diffuser shape 4mm 6mm B 125mm Z 44 b 104mm β 1 87 o B/b 1.2 β o are used in blade channels. In the gap area between inlet nozzle exit and impeller and also the area around tongue, mesh size is very smaller than that of other meshes in volute. To have a realistic velocity profile at the entrance of inlet nozzle and also to have well developed velocity profile at outlet channel, a hemisphere was added to inlet nozzle and the length of outlet channel was increased by a distance more than 10D 2. Figure2 shows a sketch of the fan with related mesh pattern and Figure3 shows mesh distribution at impeller. 2 Numerical Simulation 2.1 Fan Geometry and Mesh Generation For the present case studies, six geometries for three different inlet sizes were created by using the commercial software GAMBIT. The geometrical details are presented in Table1. A rotor with outer diameter of 216mm and 104mm width was selected such that the ratio of outer diameter to inner diameter of impeller is 0.86 as mentioned in Table1. The ratio of blade chord to impeller diameter and inlet and outlet angles of blades are selected as reported by Yahya and Patrick [9, 10]. Similarly, the other parameters viz., casing and gaps which also include ratio of casing width to impeller width and position of tongue were selected from optimum and more common ranges and values as reported by Suzuki, D.Raj, Gessner and Yan Gui [4, 11, 1, 12]. As mentioned in Table1, three sizes of inlet diameter: 175 mm (smaller than inner diameter of impeller), 190mm (nearly same with inner diameter of impeller) and 205mm (greater than inner diameter of impeller) are selected for numerical simulations. The rotational speed is selected as 900 rpm for all the case studies. To construct these geometries, unstructured tetrahedral cells were used, while the finest meshes Figure2: General view of mesh pattern Figure3: Impeller mesh distribution 2.2 Applied Numerical Procedure Due to fixed temperature in flow process, the governing equations for steady, incompressible flow are resulted from mass conservation law and ISBN:

4 momentum conservation law as given by Eqs. (1) and (2). i=1, 2, 3 Eq. (1) i=1, 2, 3 j=1, 2, 3 Eq. (2) To solve these equations and to obtain required parameters and flow pattern, a commercial CFD software, FLUENT is used. To model the turbulences the standard K-ε method with standard wall function is used and for the pressure-velocity coupling, SIMPLE algorithm is applied. For accuracy of results, second order-upwind scheme is selected for discretization of related parameters with a convergence criterion of Mass flow rate of the fluid is specified at the inlet of the nozzle and pressure outlet boundary condition is selected at the exit plane. Also Multiple Reference Frame technique is applied to model the rotating elements. 3 Experimental Validation To validate the numerical model and to compare the results, experiments were conducted such that all dimensions and parameters in both of numerical simulations and experimental setup are same as mentioned in Table1. To drive the fan, an electric motor of 0.6 kw is connected to impeller by rigid coupling to enhance the mechanical efficiency. The rotational speed in all cases was maintained constant at 900 rpm. Figure4: Fan elements and equipment design 4 Results and Discussion The results of inlet shape on efficiency at different flow rates and different inlet sizes are shown in Figures 6 and 7. The figures show an excellent match between numerical and experimental results in terms of the magnitudes as well as trends for all the six case studies considered in the present paper as shown in efficiency curves. Hence the numerical model has been validated and further studies were performed with this model. These figures show that the fans with largest diffuser inlet size of 205 mm among the other inlet diffuser sizes and nozzle inlet of 190mm among the other inlet nozzle sizes considered in the present study have the highest efficiency. In the case of using inlet diffuser, gradually by decreasing the inlet size, efficiency and maximum flow rate is decreased as shown in Figure 7, but in the case of using inlet nozzle, by increasing or decreasing the inlet nozzle size of 190mm, that is nearly same with impeller diameter, efficiency decreases and losses increases. To perform experiments and obtaining performance and efficiency curves the standard B.S. 848 [13] was followed. The arrangement of experimental setup is shown in Figure5. The performance parameters are given by flow coefficient and head coefficient as presented in Eqs. (3) and (4). Eq. (4) Eq. (3) Figure5: Variation of Efficiency with Flow Coefficient for fans with different inlet nozzle sizes ISBN:

5 Figure6: Variation of Efficiency with Flow Coefficient for fans with different inlet diffuser sizes Also by comparing efficiency curves of Figures6 and 7, it is obvious that the efficiency and performance curves for all sizes of diffuser inlets with respect to nozzle inlets are more flat around the maximum efficiency point with almost higher magnitude which conveys that fans with diffuser inlets can be operated in a wider mass flow rate band around the best performance point without significant fluctuation of efficiency. For the purpose of further analysis and finding the flow pattern, three fans with inlet nozzle sizes of 170mm, 190mm and 205mm(designated as Fan1, Fan2 and Fan3) and three fans with inlet diffuser sizes of 170mm, 190mmand 205mm(designated as Fan3, Fan4 and Fan5) with flow coefficient of φ=0.27 were considered. To study the distribution of static pressure, axial velocity and radial velocity, some contours at different axial positions are selected where they can be captured well and are presented as follows. Squirrel cage fans with forward blades are reaction type as the channel area between blades is reducing toward exit so that the fluid velocity in blade channels is increased and static pressure is reduced. Unlike other turbo machines, the static pressure of fluid is not increased in the blade passages, but it is increased in volute casing. Under such circumstances, due to high blade exit velocity, if there is an inappropriate space between the impeller and casing at the exit of impeller or faulty volute design, turbulences and losses increase significantly. To study the static pressure contours, two areas are selected as shown in Figure4. The Area1 represents the space between 30 0 plane and plane which is small space under tongue region. In this area, the contours of static pressure with almost constant radius start forming and will end in the region near outlet channel with highest cross sectional area of volute (before 90 o ). The Area2 represents the space between 30 o plane and 90 o plane. This area is selected as highest adverse pressure gradient can be observed which causes a serious problem for exit flow of corresponding blades. Figure8 represents the static pressure contours of Fan2 and Fan5 at the impeller entry. The direction of the axial coordinate is shown in Figure4. As shown in Figure7, at all axial locations in Area1 and even in space under that, the conversion of dynamic pressure into static pressure (pressure recovery) is not complete and normally the pressure of this space is as same as the pressure at blade exit. a. Fan2 at Impeller entry b. Fan2 at Z=50mm c. Fan5 at impeller entry d. Fan5 at Z=50mm Figure7: Static pressure distribution of Fan2 and Fan5 at axial position of impeller entry and Z=50mm According to Figure7 for both types of inlet, nozzle type and diffuser type, due to this adverse pressure gradient, the static pressure inside the blade channels and some parts of impeller are influenced ISBN:

6 by the adverse pressure and their pressure is increased. With respect to Figure8 it can be found that due to this adverse pressure gradient there is reverse flow that influences some part of blades and impeller inside. Also, in the space between tongue and blades at 30 o, due to narrow space, leaking of flow from tongue upper side to Area1 is not easy and there is a low amount of leakage. But leaking of flow with rotation of corresponding blades in Area1 can be the cause of intensive turbulent and vortex flow creation. Exit flow from inlet is almost axial with high magnitude reaching itself to impeller fast. In impeller whatever fluid particles move inside, gradually their axial velocity is converted to static pressure and radial velocity increases. At this condition by decreasing of Z toward the end of the impeller and with respect to Figures7 and 8, gradually the static pressure and radial velocity distribution will be more appropriate and uniform and the effect of adverse pressure gradient in Area2 and leaking flow from tongue section in Area1 will be reduced. Since the axial velocity at the impeller entry is too high, fluid particles do not have enough time to adapt themselves to impeller condition and change their direction to form radial velocity. So at the first section of impeller, blades are faced with stalling, unsteadiness and turbulences which are the causes of losses. Hence at the first sections of impeller blades, in most parts of suction side of blades there is reverse flow and corresponding turbulences and losses and finally the energy transfer is too weak at the first sections of impeller. Figure9 shows the energy transferred to the fluid particles along the blades at the angle of 180 o along the axis at different radial positions of 9cm (near blade entry) and 11.3cm (near blade exit). a. Fan2 at impeller entry b. Fan2 at Z=50mm c. Fan5 at impeller entry d. Fan5 at Z=50mm Figure9: Energy transfer along the blade width of Fan2 at the best efficiency point and at the radial positions of 9cm and 11.3cm In squirrel cage fans, static pressure and axial velocity of fluid at the inlet exit or impeller entry have a great role on flow pattern and fan efficiency. It means if at the inlet exit, axial velocity be reduced or static pressure be increased, fluid particle have enough time to reach themselves to the blade entry at the well angle and conditions, especially at the first section of impeller. Inlet shape can influence on this axil velocity and static pressure. Figure10 shows that diffuser inlet, due to relatively higher exit cross section area, have relatively higher static pressure (Figure7) and lower axial velocity at the same mass flow rate. The effect of this cause can be seen in Figure6 that by increasing the size of inlet diffuser, efficiency increases and hence losses decrease and energy transferred to along the blades increases. Figure8: Radial velocity distribution of Fan2 and Fan5 at axial position of impeller entry and Z=50mm ISBN:

7 a. Inlet nozzle b. Inlet diffuser Figure11. Types of using an inlet: as inlet nozzle or inlet diffuser In the case of Fan1 and Fan2 there is a same trend and by increasing the inlet diameter, efficiency increases. But for Fan3, since the inlet diameter is higher than the impeller diameter, exit flow from inlet nozzle, axially enters to the blade channels and destroys the flow pattern in the blades. Due to this event, losses increases and efficiency decreases. Finally, since Fan6 has the largest exit diameter among other fans, it has the highest efficiency and best performance. 5 Conclusions To find out the effect of inlet shape, inlet nozzle and inlet diffuser, at different diameters some case studies by choosing of three different inlet size with respect to impeller diameter 175mm, 190 mm and 205mm were chosen and simulations and experimental validations were performed. Results showed that there is a well matching between numerical calculations and experiments. These results showed that by increasing of inlet diffuser diameter from 175mm to 205mm, efficiency continuously is increasing, but for inlet nozzle there is same trend up to the same diameter as impeller diameter. Also efficiency curves of fans with inlet diffuser with respect to fans with inlet nozzle around the best performance point are more flatten and insignificant efficiency drop and fluctuation. The reason of difference in using and transferring the input energy in volute and blades was considered. Flow pattern in impeller showed that static pressure distribution has a great influence on components of velocity distribution. Results showed that flow at inlet exit and impeller entry is almost axially with high magnitude that by increasing of inlet diameter, static pressure is increasing while axial velocity is decreasing. By reduction of axial velocity, at the first section of impeller more number of fluid particles has enough time to reach themselves to impeller blades and form higher radial velocity magnitude. Due to this, turbulences, vortexes and related losses along the blade width will be reduced and energy transferred along the blade will be increased. Table2 represent the list of used symbols b B C D i D 1 D 2 N P Q X Y Z z α α s β 1 β 2 ρ ψ φ η μ τ Table2: List of symbol description Width of impeller Width of casing Gap between impeller and inlet Inlet diameter Inner diameter of impeller Outer diameter of impeller Rotational speed of impeller (rpm) Pressure Volume flow rate (m 3 /kg) Coordinate direction Coordinate direction Coordinate direction Number of blades Angle of tongue position Spread angle of casing Inlet angle of blade Outlet angle of blade Density(m 3 /kg) Head coefficient Flow coefficient Efficiency Viscosity of flow Stress tensor ISBN:

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