Improving performance of an industrial centrifugal blower with parallel wall volutes

Accepted Manuscript Improving performance of an industrial centrifugal blower with parallel wall volutes C. HariHaran, M. Govardhan PII: DOI: Reference:

S1359-4311(16)31391-6 http://dx.doi.org/10.1016/j.applthermaleng.2016.08.045 ATE 8844

To appear in:

Applied Thermal Engineering

Received Date: Revised Date: Accepted Date:

17 March 2016 19 July 2016 7 August 2016

Please cite this article as: C. HariHaran, M. Govardhan, Improving performance of an industrial centrifugal blower with parallel wall volutes, Applied Thermal Engineering (2016), doi: http://dx.doi.org/10.1016/j.applthermaleng. 2016.08.045

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IMPROVING PERFORMANCE OF AN INDUSTRIAL CENTRIFUGAL BLOWER WITH PARALLEL WALL VOLUTES

Corresponding author: C. HariHaran*, Thermal Turbomachines Laboratory, Department of Mechanical Engineering, Indian Institute of Technology Madras, Chennai, India - 600036 Mail ID: [email protected],

M. Govardhan Thermal Turbomachines Laboratory, Department of Mechanical Engineering, Indian Institute of Technology Madras, Chennai, India - 600036 Mail ID: [email protected]

IMPROVING PERFORMANCE OF AN INDUSTRIAL CENTRIFUGAL BLOWER WITH PARALLEL WALL VOLUTES C. HariHarana*, M. Govardhanb Thermal Turbomachines Laboratory, Department of Mechanical Engineering, Indian Institute of Technology Madras, Chennai, India [email protected] a, [email protected]

ABSTRACT In this article, parallel wall volutes are proposed as an alternate and energy efficient volute to the commonly used rectangular volute for an industrial centrifugal blower. A detailed comparison of performance between parallel wall and rectangular volute are made for four different width ratios. The stage performance analysis suggests that the parallel wall volutes perform better in terms of both specific work developed and total-total isentropic efficiency. From component performance assessment, parallel wall volutes found to have higher static pressure recovery and lower loss for the full operating range considered. The flow field reveals that the parallel wall volute shows more uniform static pressure distribution at the volute inlet compared with the rectangular volute. The detailed aerodynamic performance analysis reveals that the overall performance of a centrifugal blower can be improved up to 6% with parallel wall volute. Moreover, it is demonstrated that the rectangular volute can be replaced with parallel wall volute without modifying foundation, inlet and outlet duct.

Keywords: centrifugal blower, volute aerodynamics, Parallel wall volute, Rectangular volute, efficiency improvement

NOMENCLATURE A

- cross sectional area (m2)

Aθ

- area at impeller exit (m2)

Cp

- pressure recovery coefficient

Cω

- volute loss flow coefficient

Ip

- input power to impeller (W)

m

- mass flow rate (kg/s)

P

- static pressure (N/m2)

P0

- total pressure (N/m2)

PU

-

R/R2

- ratio of radial postion to impeller radius

T0

- total temperature (K)

U2

- impeller tip velocity (m/s)

w

- isentropic specific work (m2/s2)

Z/B

- ratio of axial position to blade height

α2

- flow angle at impeller exit (deg.)

η

- total to total isentropic efficiency of stage (%)

Ψ

- head coefficient

impeller dynamic pressure (N/m2)

Subscripts 0

- inlet of suction duct

1

- exit of suction duct

2

- inlet of impeller

3

- exit of impeller

4

- inlet to volute

5

- exit of volute

d

- design value

Abbreviations B2

- impeller exit width (m)

B5

- volute width (m)

R 4.0

- Ratio of volute width to impeller width = 4.0 of parallel wall volute

R 5.0

- Ratio of volute width to impeller width = 5.0 of parallel wall volute

RE 4.0 - Ratio of volute width to impeller width = 4.0 of rectangular volute RE 5.0 - Ratio of volute width to impeller width = 5.0 of rectangular volute

1. INTRODUCTION Volute is one of the key components in a centrifugal machine, its function is being to collect the fluid from impeller and transfer to the delivery pipe. The main parameters affecting the volute performance are tongue clearance, shape and cross sectional area of the volute [1], [2]. The modern design philosophy of industrial turbomachines is to reduce the cost by making more efficient and compact machine. In order to design such a machine, there is an increase in trend towards (i) setting a higher ratio between volute width (B5) and impeller width (B2), and (ii) partial pressure recovery in the volute. This strategy has been proven to have a profound effect on the performance, especially at off design mass flow rates as reported by Hariharan and Govardhan [3]. However for full operating range, there is a need to understand the effects of shape and area ratio towards designing a highly efficient volute. The widely used industrial cross sections shapes of the volute are parallel wall, rectangular and circular. In general, there is a scarcity of literature available for the effects of shape and area ratio of volute. A few research works have been reported in this direction, mostly for rectangular and circular volutes. The collective works are summarized as follows. Yang et al. [4] numerically studied the effect of various cross sectional shape on centrifugal pump volute (round, rectangular, trapezoid and horseshoe shaped), and they found that all volutes produce almost same head and efficiency at design flow rate. Reunanen [5] performed numerical simulation and experimentally evaluated the effect of cross section shape on the volute performance. They considered the rectangular and semicircular cross sections and they revealed that at low mass flow rates the change in shape has influence on both pressure recovery and efficiency. Jin et al. [6] conducted the numerical study on the effect of change in cross sectional area on the performance of centrifugal pump. Whitfield and Johnson [7] experimentally assessed the square volute of turbo charger for variable cross sectional area conditions. They found that an optimum ratio volute improves the performance for full operating range. Flow field inside the volute has been investigated by many researchers, and the salient works are as follows. Hagelstein et al. [8] conducted both numerical and experimental studies on rectangular volute. They presented the flow field, pressure recovery and total pressure loss coefficient along different cross section of the volute. Dilin et al. [9] carried out experiments for exploring the effect of tongue on the flow structure and loss mechanism of two different rectangular volutes. Van den Braembussche et al. [10] demonstrated the theoretical model to

predict the pressure loss and pressure recovery coefficient at design and off design mass flow rates for a rectangular volute. Ayder et al. [11] employed Euler solver and numerically simulated the flow inside the circular volute, incorporating loss model for frictional and shear losses they are able to predict the pressure distribution along the volute. Hagelstein et al. [12] experimentally studied the effect of non-uniform pressure distribution created by concentric volute and its effect on the upstream component. Hillewaert and Van den Braembussche [13] presented interface model to simulate the interaction between impeller and volute and, by iteratively updating the boundary values at the interface they predicted the pressure distribution along the volute. Based on the comprehensive survey, it may be concluded that most of the studies are on the compressors, whereas the parallel wall volutes are preferred in blowers. Thus, only seldom research on parallel wall volte has been reported in open literature. Recently, Hariharan and Govardhan [3] numerically studied the parallel wall volute for variable width ratios and they reported that the higher ratio volutes outperform especially at lower mass flow rates. Heo et al. [14] conducted the optimization studies on impeller blades of an centrifugal blower with parallel wall volute. Baloni [15] conducted optimization studies on centrifugal blower volute using taguchi method and they concluded that area ratio has higher influence on pressure nonuniformity at impeller outlet. Xu and Mao [16] numerically evaluated the aerodynamic performance and acoustics characteristics of the centrifugal fan with parallel wall volute. In recent years many researchers like, Viholainen et al. [17], Degueurce and Tersiguel [18], Al-busaidi and Pilidis [19] pointed out the effect of centrifugal blowers’ energy efficiency on the environment and cost. Volute is known to be the least performing component in centrifugal blowers [20], [21]. To improve the blower performance volute would be the ideal place to start. However, replacing the existing volute in industrial environment is extremely difficult due to (i) modifying the foundation (ii) rearranging the inlet and outlet duct (iii) additional floor area requirement. But, the rectangular and parallel wall volutes have an identical external dimension, which give the unique advantage of interchanging without any difficulties. There is no open literature available, comparing the performance of parallel wall volute and its equivalent rectangular volute. Thus, the present study explores numerically the effect of shape and area ratio extensively at both design and off design mass flow rates for four different parallel wall volutes and their equivalent rectangular volutes. The better volute has been suggested by

assessing the aerodynamic performance both stage and component level like, operating range, better static pressure recovery and low total pressure loss. 2. DESIGN AND GEOMETRY DETAILS The blower is designed with 12 circular arc blades and runs at 3000 rpm delivering 28 kg/s mass flow rate and 24000 m2/s2 specific work. The blower has an operating range between 20 kg/s to 32 kg/s. The blower essentially consists of impeller, inlet duct and volute. The impeller and inlet duct are the same one used in earlier [22], the geometrical details of the impeller and inlet duct are presented in Table 1. The four different parallel wall volutes are designed based on constant angular momentum principle for width ratios of 3.0, 3.5, 4.0 and 5.0, designated as R 3.0, R 3.5, R 4.0 and R 5.0, respectively. The equivalent rectangular volutes have been designed by keeping all external dimensions same as parallel wall volute and providing separator wall as shown in Fig.1. The cross sectional area for parallel wall volute includes area (I) and area (II), whereas rectangular volutes have only area (I). In order to study the effect of shape and area ratio alone, all other parameters like tongue gap, tongue position with respect to volute exit and exit cone dimensions are kept constant. The rectangular volutes designated as RE 5.0, RE 4.0, RE 3.5 and RE 3.0 is equivalent to that of R 5.0, R4.0, R 3.5 and R 3.0 parallel wall volutes, respectively. Figure.2 shows the ratio between cross sectional area and impeller exit area at different peripheral angles. The horizontal axis represents the circumferential position of the cross section (θ). An angle 32º and 360º are the tongue and throat position, respectively. Although, the tongue gap is same for all the volutes, the parallel wall volutes have a higher area compare to rectangular volutes. This is because of the change in volute height and the inclusion of area (II) and the same reason can be attributed for the entire volute. In parallel wall volute the ratio 5.0 volute has highest area for almost all the cross sections compare to other volute, besides the area follows increasing trend towards increasing width ratio. On the other hand, in rectangular volute, the area follows decreasing trend with increasing width ratio, except at tongue the ratio 5.0 volute has a slightly more area. This trend is due to the logarithmic increment and constant tongue clearance in all the volutes.

3. COMPUTATIONAL METHODOLOGY The numerical simulations are done using commercial CFD code ANSYS CFX 14. To simulate the pressure disturbance due to asymmetry shape of the volute and velocity profile at the impeller inlet. The computational domain includes inlet duct, full 3-D volute and impeller. A structured hexahedral mesh is generated using ANSYS ICEMCFD 14. To reduce numerical diffusion during simulation, high quality mesh is generated at all domains by maintaining minimum angle between diagonals 18º and volume expansion factor of 30 [23]. The steady sate simulations are carried out for five different mass flow rates at design and off design points for each volute. The continuity, energy and Navier-stokes equation are solved using second order discretization technique. The high speed turbulent flow inside the volute is captured using k-ɛ turbulence model. Due to compressible nature of the flow air ideal gas assumption is included. Total pressure inlet and mass flow outlet boundary condition are imposed. The interface between impeller and volute are facilitated by frozen rotor technique. All the walls are considered as adiabatic and the required y+ value of 30 for k-ɛ turbulence model is maintained. The convergence criteria for simulation are sets at fifth order of residuals, as well as the key variables like pressure at outlet, mass flow at inlet are monitored. The simulations are continued until the steady state is attained. At lower mass flow rate the variables continue to oscillate up to larger number of iterations, especially for rectangular volutes. At lowest mass flow rate and with RE 3.0, RE 3.5 and RE 4.0 volutes, the numerical oscillations continued after several thousand iterations due to unsteady nature of the flow physics. For the present investigation such a point is removed from calculation. Figure.3 shows full computational domain of the blower of RE 5.0 volute, and the different stations are also marked. The grid independency is ensured using Richardson grid independence study technique [25] for all volutes. For the sake of brevity only the R 5.0 grid independence study results are presented Table 2. The computational methodology is validated using test rig data reported in Bhargava [24]. The tested blower consists of inlet duct, impeller, parallel wall volute and discharge duct with throttle valve, the tested and design blowers has almost same nondimensional shape number. The geometrical details of test rig data are provided in table 3. The aerodynamic performance of the blower, both experimental and computational is presented in table 4. The detailed comparison of static pressure profile across the volute cross section is shown in Fig.4. Both the experimental and CFD results are matching closely.

4. RESULTS AND DISCUSSION Numerical simulations are performed at five different mass flow rates for each of the volutes namely, R 5.0, R 4.0, R 3.5, R 3.0, RE 5.0, RE 4.0, RE 3.5 and RE 3.0. The results are discussed based on the stage performance and volute performances. Flow field inside the blower is visualized using contours and vectors. 4.1 Stage head coefficient and isentropic total-total efficiency The stage head coefficient is defined as the ratio between stage specific work and square of impeller tip velocity. (1) Where,

(2)

Figure.5 (a) shows the head coefficient plotted for different mass flow rates. Even though the impeller geometry and mass flow rates are same, depending on the volute, the specific work developed by each machine varies. The parallel wall volute outperforms the rectangular volute at all mass flow rates. At the design mass flow rate all the volute develop almost same specific work. At off design mass flow rates especially at lower mass flow rate, the rectangular volutes shows drop in specific work, a typical values at BEP (best efficiency point) is about 6% less. The stable operating range of machine also gets reduced for rectangular volute. Especially, for RE 3.5 and R 3.0 volutes, the lowest feasible mass flow is found to be only m/md=0.9 and 1.0, respectively. The total-total isentropic stage efficiency is defined as the ratio between useful work transferred to the fluid and shaft input power. (3) It can be inferred from Fig.5 (b) that the total-total isentropic efficiency is high for the parallel wall volute for the full operating rage. Further, the difference in efficiency between rectangular and parallel wall volute is not as vigorous as specific work at BEP which is about only 1.5%. Similarly, the input power to impeller also gets reduced to considerable extent, although it is not severe as specific work. Thus overall drop in isentropic efficiency can be attributed to drop in both the specific work and input power.

In the parallel wall volute, near the design mass flow rate, R 3.5 and R 3.0 volutes performs better both in terms of specific work and efficiency, whereas at lower mass flow rate R 5.0 and R 4.0 are found to be better. In the rectangular volute, at design and m/md=1.1 mass flow rate RE 3.0 and RE 3.5 perform better and also follows same trend as that of parallel wall volute. Lower than design mass flow rate (m/md < 1), RE 5.0 and RE 4.0 performs almost same. Whereas, in RE 3.5 and RE 3.0 flow becomes highly transient, thus the steady state comparison could not be viable. To identify a better volute for full operating range, further investigations on component level and flow physics only the ratio 4.0 and ratio 5.0 volutes are analysed. 4.2 Pressure recovery and Pressure loss coefficients Pressure recovery coefficient (Cp) is defined as the ratio between the static pressure recovered in the volute to the dynamic pressure at the impeller exit. (4) Figure.6 shows the static pressure recovered for different mass flow rates. Both the parallel and rectangular volutes of Ratio 4.0 have higher pressure recovery compared to ratio 5.0 volutes. It is interesting to note that the stage performance of R 5.0 is better than RE 4.0; yet, the volute performance gets degraded by 2% drop in pressure recovery. The parallel wall volutes perform better by 4% higher pressure recovery compared to the equivalent rectangular volutes. The results exhibit pressure recovery is more influenced by the width ratio than the shape of volute. Loss coefficient (ω) is defined as the ratio between the total pressure losses in the volute to the dynamic pressure at the impeller exit. (5) The total pressure loss includes all the losses in the volute: meridional velocity dump loss, frictional loss, tangential velocity dump loss and exit cone loss [26]. Rectangular volutes produce higher total pressure loss compared to parallel wall volutes. It is evident from Figure.5 that the pressure loss is directly proportional to width ratio. Moreover, the difference in loss between higher and smaller ratio of rectangular volute is very small where compared to the difference in loss for the equivalent parallel wall volutes. The RE 4.0 volute has higher total pressure loss compared to R 5.0 volute despite its high Cp value.

4.3 Flow field at volute inlet The volutes are designed based on the constant angular momentum principle. The assumptions made are flow at the impeller exit is uniform and the fluid follows a logarithmic flow path. In the present design, there is no diffuser after impeller, so the fluid from the impeller exit is directly delivered to the volute. The average pressure along the volute keeps increasing, this variation in pressure is transferred to the upstream components and it leads each impeller blade to get exposed to different exit (throttle) conditions. This essentially changes the impeller and volute performance. Static pressure at volute inlet: To better understand the influence of volute on the impeller performance, the static pressure and flow angle are plotted at the volute inlet. The flow variables are averaged both in circumference and span wise for a sector of 30º. The detailed investigations are carried out at design mass flow rate, which is closer to BEP. Figure.7 (a) show averaged static pressure along the circumference of volute inlet for a specific case m/md =1.0. The pressure along the circumferential direction increases from tongue to the volute exit for all the four cases, however the pressure fluctuates at the volute inlet. It may be mentioned here that tongue is the geometric location (32º) where in the flow is divided into recirculating flow and discharge flow. In the parallel wall volute, pressure is almost uniform along the circumference of the volute inlet, except for a small drop noticed between 225º to 330º. At other sector angles, pressure is almost same for both the volutes. In the rectangular volute, along the circumference pressure decreases from the tongue and it reaches minimum value at 105º. This could be the effect of a smaller cross sectional area near the tongue (see Fig.2). Thereafter, pressure increases and reaches a maximum value just before the tongue 15º (throat), where the average pressure across the volute cross section is maximum. The ratio 4.0 volute has slightly higher pressure along the volute inlet compare to ratio 5.0 volute. However, the influence of width ratio is relatively small compared to influence of the shape. Figure.8 shows typical static pressure contours at the volute inlet for a specific case m/md =1.0. As concluded from earlier discussion, the width ratio has relatively smaller influence on pressure distribution than shape of the volute, thereby only ratio 5.0 contours are presented. Along the circumference pressure decreases from hub to shroud for both the volutes, however this decreasing pressure trend varies with respect to the type of volute. Typically, the decreasing

pressure trend is noticed from 0 to 180º in parallel wall volute, whereas in the case of rectangular volute this trend is confined to 150º. In this region, for the rectangular volute pressure is relatively low and its gradient is high for any given angle. For a particular blade passage the pressure decreases from pressure side to suction side, this decreasing trend is present between 30º to 240º in parallel wall volute, whereas in the rectangular volute it is prominent in 240º to 360º sector. In both the volute highest (I) and lowest (II) pressure are noticed just before the tongue (15º) and at the tongue portion of the shroud, respectively. The intensity and area of these zones are prominent in rectangular volute. This could be due to the dominant recirculating flow in the vicinity of tongue. The blade wake (III) is quite evident in all the volutes, except in the vicinity of tongue. It can be inferred that parallel wall volute show more uniform pressure compare to rectangular volutes, this implies that parallel wall volute helps in reducing the structural load on the impeller. Flow angle and velocity at volute inlet: Figure.7 (b) shows averaged flow angle along the circumference of volute inlet for a specific case m/md =1.0. Flow angle is one of the key variable in volute design, while designing it is assumed constant along the circumference (32º). However, Fig.7 (b) contradicts this assumption, where flow angle varies inversely with pressure due to change in the meridional velocity. Besides, it is also influenced by the formation of swirl inside the volute. Parallel wall volute has almost uniform flow angle along the circumference, closer to design value. Except a small increase in flow angle closer to 345º, this could be due to the interaction between swirl inside the volute and impeller exit flow. In rectangular volute, flow angle increases across the tongue (30º); which could be the result of increased meridional velocity. This trend continues up to 75º, thereafter, flow angle becomes almost constant till 135º. From there, it decreases and reaches a minimum value at 15º, wherein the static pressure is maximum. The present design is non pre whirl design; hence the circumferential velocity represents the work done by the impeller. Conversion of circumferential velocity in the volute is the main contributor for the static pressure recovery. Figure.9 debates the variation of circumferential velocity at the volute inlet for specific case m/md =1.0. In the vicinity of tongue, high velocity (I) is noticed in the suction side and it occupies about (5%) of the passage. This high velocity zone grows along the circumference up to 270º; afterwards it decreases. The reason could be flow turning from pressure side (PS) to suction side (SS) due to the blade wake. In the span wise

direction high velocity is noticed above an axial distance of 80% from the hub. This could be due to secondary flow, the fluid entering the pressure side at the blade leading edge and while moving in the meridional path, the flow deviates and occupies this region (top 20% axial distance). In the rectangular volute from tongue to 210 º high velocity is observed near suction side of the shroud; it occupies about 3% of the passage. In the vicinity of tongue this high velocity present at middle of the passage along the circumference, move towards suction side while it also grows. At 330 º this high velocity spread across the entire suction surface and occupies almost 20% of the passage. A zone of low velocity is noticed closer to hub at the middle of the passage irrespective of angle and it decrease along the circumference. This zone occupies almost 50% of the passage near the tongue, whereas at 210º it occupies only 5% of the passage. This change in velocity profile is due to the variation of static pressure and slip factor. In overall, the velocity profile in parallel wall volute is relatively uniform along the circumference compare to rectangular volute. Meridional velocity at volute inlet represents the mass flow distribution and localized throttle condition. Figure.10 shows a typical meridional velocity contour at volute inlet for a specific case m/md =1.0. In parallel wall volute, near the tongue, a zone of low velocity (I) is noticed at 50% span of suction surface and it grows along the circumference, especially between 180 º and 270 º this growth is prominent. In the rectangular volute near the tongue a small zone of low velocity (I) is noticed at middle of the passage, along the circumference it move towards suction surface while, it also grows. At 180 º it reaches suction surface and occupies top 20% span of suction surface, whereas at 330 º it occupies almost 30% of passage from suction surface. This variation in velocity profile is predominately due to (i) the change in static pressure along the circumference of volute (ii) flow separation and circulation in the impeller. 4.4 Flow field at volute cross section Velocity profile at the volute cross-section: The velocity inside the volute can be split into two components; cross flow velocity (velocity parallel to cross section) and through flow velocity (velocity perpendicular to cross section). Dissipation of the cross flow velocity is the main contributor for volute losses (meridional dump loss). Figure.11 shows the cross flow velocity contour super imposed with velocity vector at two different cross sectional planes 90º and 270º (see Fig.3) for ratio 5.0 volutes at m/md =1.0. For all the cross sections, the maximum cross flow velocity is noticed at the volute inlet and it decreases

while moving towards side wall of the volute. At the 90º cross sectional plane in parallel wall volute, high velocity is noticed at two regions. (i) Closer to volute inlet, this section is filled with fluid entering from impeller into the volute (new fluid). (ii) Closer to top wall of the volute. The fluid entering closer to tongue move circumferentially and before reaching the 90º cross section it is pulled towards the top wall, because of the pressure difference between the inner and outer radius. In volute the pressure keeps increasing towards higher radius this force is balanced by centrifugal and coriolis forces. The fluid between shroud and top wall (I) has low pressure and kinetic energy. Whereas, the fluid coming from the impeller has high pressure and kinetic energy, which leads to the fluid moving towards top wall of the volute. The remaining portion of this cross section has low velocity, since major portion of the cross section is occupied by recirculating flow. Rectangular volute has a higher velocity compared to parallel wall volute across the entire cross section. This could be due to smaller cross sectional area (see Fig.2). High velocity is noticed near volute inlet same as that of parallel wall volute. Besides, a small zone of high velocity is present near side wall. The reason could be the acceleration of recirculating flow coming from tongue due to the presence of low pressure zone. At the 270º cross sectional plane the cross flow velocity profile is almost same for rectangular and parallel wall volutes for the equivalent radius. A high velocity zone is seen closer to volute inlet, which is similar to that of 90º plane. In addition a small zone of high velocity is observed closer to top wall of the volute. The size and influence of high velocity zone is more for rectangular volutes. The velocity between shroud and top wall is low, as this region is mostly occupied by recirculating flow. Inferred from the velocity vectors, in both the volutes cross flow velocity forms a single vortex profile irrespective of the cross sections. The center of vortex is occupied by recirculating fluid and it wrapped by a new fluid. In parallel wall volute only part of the recirculating flow is covered by new fluid, whereas in the rectangular volute entire recirculating flow is covered by a new fluid. The velocity is minimum at the vortex center and it increases with radius (forced vortex); this increment is relatively small in parallel wall volute. At 90 º cross section in the parallel wall volute, the center of vortex is noticed closer to top wall, whereas in the rectangular volute it is placed at cross sectional center. At 270º in both the volutes vortex center is present at cross sectional center.

Figure.12 shows the through flow velocity at two different cross sectional planes 90º and 270º of the ratio 5.0 volutes for a specific case m/md =1.0. At 90º, in parallel wall volute low through flow velocity is noticed between inlet duct wall and R ≈ 1.0, due to low energy recirculating fluid occupied this region. Wherein, in the radial direction velocity is almost linearly increases from inlet duct to R ≈ 1.0, in the axial direction velocity increases from shroud to top wall, however, this increment is relatively small closer to shroud. Above R ≈ 1.0 velocity is almost uniform and high. In the rectangular volute entire cross section has almost uniform and higher velocity, except for a small zone of low velocity closer to side wall. This could be due to the fact cross sectional area of rectangular volute is smaller than parallel wall volute (see Fig.2) and the mass flowrate is almost same. At 270º, parallel wall volute has a unique velocity profile. The velocity increases from inlet duct to R ≈ 1.0 (I), and maintains almost uniform velocity between R ≈ 1.0 and R ≈ 1.4 (II), thereafter it decreases up to side wall (III). The forces in fluid element are centrifugal force (F c), Coriolis force (Fco) and pressure force (Fp). And the available energy (E) in the fluid elements is present in the form of (i) pressure and (ii) momentum energy. Figure.13 shows the major force along the radial direction in the fluid element (A) of a parallel wall volute. In region (I), the available energy (E1) is low due to the presence of recirculating fluid, but it linearly increases with radius. To balance the high pressure force from neighboring (higher radius) fluid element, the fluid decelerates and partially converts its available energy into pressure energy. Therefore, velocity linearly increases in this region. In region (II), the available energy (E 2) is high, because this region is mostly occupied by new fluid coming from the impeller. To balance the pressure difference the lower radius fluid (I), fluid accelerate and partially convert its available energy into momentum energy. Thereby, velocity is high in this region. In region (III), the available energy E3 is almost same as that of E2. With the increase in radius, fluid is decelerating (free vortex) and pressure increases, this balances the higher centrifugal force in the neighboring fluid (lower radius). Thereby, velocity is linearly decreasing in this region. In the rectangular volte also the velocity profile and values (region II and III) are almost same as that of parallel wall volute for the equivalent radius. Despite the cross sectional area of rectangular volute being less than parallel wall volute for the same mass flow rate, the velocity is almost the same. This could be due to (i) most of the additional area (II) (see Fig.1) in parallel

wall volute participates very little to mass flow rate and (ii) in parallel wall volute the additional recirculating fluid entered through area (II) at the tongue. Total pressure distribution at the volute cross-section: Figure.14 shows the total pressure distribution at two different cross sectional planes 90º and 270º for ratio 5.0 volute for a specific case m/md =1.0. At the 90º cross sectional plane, the total pressure linearly increases from inlet duct wall to side wall for the parallel wall volute. The region near volute inlet to side wall is dominated by the fluid coming from impeller having high total pressure. The low total pressure is identified between shroud and top wall, where the flow is dominated by recirculating flow. This phenomenon in the parallel wall volute is due to the movement of high pressure fluid from outer radius towards inner radius. While recirculating, the fluid loses its energy due to turbulence and mixing. On the other hand the flow from volute inlet move towards top wall during which static pressure is increased and the velocity gets decreased. This high pressure fluid encounters low pressure low energy fluid in the region between shroud and top wall, which accelerates and mixes (and loose its energy) then it is pushed back closer to the volute inlet. This cycle continue till it reaches the tongue. At the same cross sectional plane the rectangular volute also shows higher total pressure closer to side wall same as that of parallel wall volute. A small low pressure zone is observed near separator wall, since this zone is occupied by recirculating flow. Compared to parallel wall volute, the recirculating flow has a higher loss in total pressure between tongue and 90º cross sectional plane. At 270º cross sectional plane low pressure zone (I) is noticed at the center for rectangular volute, which is occupied by recirculating flow. At the same cross sectional plane, parallel wall volute also has low pressure zone at the center, besides the presence of low pressure zone in between shroud and top wall. Static pressure recovery at the volute cross-section: Figure.15 shows the static pressure recovery coefficient (Cp) at two different cross sectional planes 90º and 270º of ratio 5.0 volutes for a specific case m/md =1.0. At 90º, in the parallel wall volute, Cp is negative above the shroud between inlet duct wall and R = 1.0 (volute inlet radius). This could be the result of higher total pressure loss in this region. In this region Cp is linearly increasing from inlet duct wall and it is almost uniform along the axial direction. Above, R = 1.0 the Cp becomes positive due to the recovery of static pressure from deceleration of circumferential velocity. In this region Cp profile follows high value closer to both top wall and bottom wall. In rectangular volute Cp is negative for a small zone closer to separator wall at Z/B2

= 2.5. This zone is mostly occupied by low energy recirculating fluid; in addition flow is accelerating due to the pressure difference and curvature of the separator wall. Near the side wall high value Cp is noticed, especially in the (Z/B2 < 1.0) and (Z/B2 > 4.0) region. At 270º, in the parallel wall volute between inlet duct wall and R = 1.0 Cp is negative similar to that of 90º. Above this radius (R = 1.0), Cp is almost linearly increasing with radius and constant along the axial direction, due to the recovery of static pressure from the through flow velocity deceleration. Also in the rectangular volute Cp is linearly increasing from volute inlet to side wall same as that of parallel wall volute, except a small drop in Cp closer to cross sectional center. This may be the result of strong vortex formation at cross sectional center (see Fig.11 b) However, the Cp values are slightly less compared to the parallel wall volute at the equivalent radius. 5. CONCLUSION The effects of shape and cross sectional area have been explored numerically for four different area ratio parallel wall volute and their equivalent rectangular volutes, and the salient points are listed here. 

Regarding stage, parallel wall volutes outperform the rectangular volutes, both in terms of isentropic specific work and isentropic total-total efficiency by 6% and 2%, respectively. Also, the stable operating range of the rectangular volute is significantly small, especially for smaller width ratio volutes.



The parallel wall volutes exhibit higher static pressure recovery compared to the equivalent rectangular volutes. The pressure recovery is more influenced by the volute width ratio than the volute cross sectional shape.



The rectangular volutes produce more loss compared to the parallel wall volutes, and the loss coefficient is found to be directly proportional to the width ratio.



Parallel wall volutes have more uniform static pressure at the volute inlet, which could essentially reduce the structural load on bearings.

Thus, for the same external dimensions, the parallel wall volutes have almost 40% more cross sectional area, yet compact, more economical and energy efficient for the full operating range. The overall investigation reveals that the parallel wall would be an ideal replacement for the existing low efficient rectangular volutes without modifying foundation or refabricating any component like inlet/ outlet duct.

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LIST OF FIGURES Figure 1 Comparison of R 5.0 and RE 5.0 volute (a) cross sectional view (b) isometric view Figure 2 Cross section area at different circumferential position Figure 3 3-D view of fan assembly for RE 5.0 volute Figure 4 Static pressure across the volute width Figure 5 stage performance for different shape of the volutes: (a) head coefficient, (b) isentropic total-total efficiency Figure 6 Pressure recovery and loss coefficient for different shape of the volutes Figure 7 (a) Averaged static pressure, (b) Flowangle at volute inlet for m/md=1.0 Figure 8 Circumfrential directional variation of static pressure at the volute inlet for m/m d=1.0: (a) R 5.0 (b) RE 5.0 Figure 9 Circumfrential directional variation of circumferential velocity at at the volute inlet for m/m d=1.0: (a) R 5.0 (b) RE 5.0 Figure 10 Circumfrential directional variation of meridional velocity at the volute inlet for m/md=1.0: (a) R 5.0 (b) RE 5.0 Figure 11 Cross flow velocity contours at various cross sections of the ratio 5.0 volute: (a) R 5.0 (b) RE 5.0 (left column 90º, right column 270º) Figure 12 Through flow velocity contours at various cross sections of the ratio 5.0 volute: (a) R 5.0 (b) RE 5.0 (left column 90º, right column 270º) Figure 13 Forces along the radial direction on the fluid element in the parallel wall volute Figure.14 Total pressure contours at various cross sections of the ratio 5.0 volute: (a) R 5.0 (b) RE 5.0 (left column 90º, right column 270º) Figure 15 Static pressure recovery coefficient contours at various cross sections of the ratio 5.0 volute: (a) R 5.0 (b) RE 5.0 (left column 90º, right column 270º)

LIST OF TABLES Table 1 Design and geometrical values of impeller and inlet duct Table 2 Richardson grid independence study Table 3 Geometrical details of test rig Table 4 Validation data

(a)

separator wall tongue

inlet duct

top wall

Parallel wall Rectangular

side wall

(b) Figure 1 Comparison of R 5.0 and RE 5.0 volute (a) cross sectional view (b) isometric view

1.5 R 5.0 R 4.0 R 3.5 R 3.0

1.2

RE 5.0 RE 4.0 RE 3.5 RE 3.0

A/A

0.9

0.6

0.3

0.0 0

60

120



180

240

300

360

Figure 2 Cross section area at different circumferential position

Impeller Inlet duct 90º

5

1

0

3 270º

Volute

Volute inlet

Figure 3 3-D view of fan assembly for RE 5.0 volute *0-Inlet, 1-Impeller inlet, 3-Volute inlet and 5-Outlet 0.8

CFD Bhargava (1978)

static pressure

P3/Pu

0.6

0.4

0.2

0.0 0.0

0.5

1.0

Normalized axial distance (Z/B2)

1.5

Figure 4 Static pressure across the volute width

2.0

0.50

head coefficient ()

0.45

0.40

0.35

R 5.0 R 4.0 R 3.5 R 3.0

RE 5.0 RE 4.0 RE 3.5 RE 3.0

0.30

0.7

0.8

0.9

1.0

1.1

1.0

1.1

normalized mass flow rate (m/md) (a)

total-total efficiency ()

(%)

80

75

70

65

R 5.0 R 4.0 R 3.5 R 3.0

RE 5.0 RE 4.0 RE 3.5 RE 3.0

60

0.7

0.8

0.9

normalized mass flow rate (m/md)

(b) Figure 5 stage performance for different shape of the volutes: (a) head coefficient, (b) isentropic total-total efficiency

0.6

C

0.5 0.5

0.4

CP RE 5.0

C RE 5.0

CP RE 4.0

C RE 4.0

CP R 5.0

C R 5.0

CP R 4.0

C R 4.0

0.4

0.3

0.3 0.7

0.8

0.9

loss coefficient

pressure recovery coefficient

CP

0.6

0.2 1.1

1.0

normalized mass flow rate (m/md)

Figure 6 Pressure recovery and loss coefficient for different shape of the volutes

R 5.0 R 4.0 RE 5.0 RE 4.0 Tongue

90 0.6

120

60

90 36

24

30

150 0.2

12

180

0

0.2 330

210 0.4

6 0

180

0

6 12 18

330

210

24 30

0.6

R 5.0 R 4.0 RE 5.0 RE 4.0 Tongue

30

150

18

flow angle 2

static pressure P3/PU

60

30

0.4

0.0

120

240

300 270

36

240

300 270

(a) (b) Figure 7 (a) Averaged static pressure, (b) Flowangle at volute inlet for m/md=1.0

S

T

Normalized axial distance (Z/B2)

I (a)

PS

III

II

SS

H

I (b)

II

III Circumferential angle (θ) (deg) (N/m2)

* (T) tongue, (S) shroud, (H) hub, (PS) pressure side, (SS) suction side Figure 8 Circumfrential directional variation of static pressure at the volute inlet for m/md=1.0: (a) R 5.0 (b) RE 5.0

Normalized axial distance (Z/B2)

I (a)

I (b)

Circumferential angle (θ) (deg) (m/s)

Normalized axial distance (Z/B2)

Figure 9 Circumfrential directional variation of circumferential velocity at at the volute inlet for m/md=1.0: (a) R 5.0 (b) RE 5.0

I

(a)

(b)

I

Circumferential angle (θ) (deg) (m/s)

Figure 10 Circumfrential directional variation of meridional velocity at the volute inlet for m/md=1.0: (a) R 5.0 (b) RE 5.0

TW

Normalized axial distance (Z/B2)

(a)

I

IW

SW

I

S H

BW

TW

SEW

(b) SW I BW

Normalized radial distance (R/R2) (H) hub, (S) shroud, (I) volute inlet, (BW) bottom wall, (IW) inlet duct wall, (TW) top wall, (SEW) separator wall, (SW) side wall Figure 11 Cross flow velocity contours at various cross sections of the ratio 5.0 volute: (a) R 5.0 (b) RE 5.0 (left column 90º, right column 270º)

A II

I

III

Normalized axial distance (Z/B2)

(a)

(m/s)

(b)

Normalized radial distance (R/R2) Figure 12 Through flow velocity contours at various cross sections of the ratio 5.0 volute: (a) R 5.0 (b) RE 5.0 (left column 90º, right column 270º)

I

III

II Fc

Fc Fp

Fp

Fc

Fc Fp

Fp

(Fc) centrifugal force, (Fp) pressure force Figure 13 Forces along the radial direction on the fluid element in the parallel wall volute

I

Normalized axial distance (Z/B2)

(a)

2

(N/m )

(b)

I

Normalized radial distance (R/R2) Figure.14 Total pressure contours at various cross sections of the ratio 5.0 volute: (a) R 5.0 (b) RE 5.0 (left column 90º, right column 270º)

Normalized axial distance (Z/B2)

(a)

(b)

Normalized radial distance (R/R2) Figure 15 Static pressure recovery coefficient contours at various cross sections of the ratio 5.0 volute: (a) R 5.0 (b) RE 5.0 (left column 90º, right column 270º)

LIST OF TABLES Table 1 Design and geometrical values of impeller and inlet duct Impeller Inlet Diameter

0.6 m

Impeller Exit Diameter

1.3 m

Inlet Blade angle

31°

Exit Blade angle

48°

Number of blades

15

Inlet duct Diameter

0.8 m

Radius of curvature of Inlet duct

0.94 m

Table 2 Richardson grid independence study [22] No of Elements N1, N2, N3

9604524, 4196409, 2109898

Grid refinement factor (r21 )

1.32

Grid refinement factor (r32)

1.31

Critical variable value for grid 1(φ1)

18160.1

Critical variable value for grid 2 (φ2)

18244.83

Critical variable value for grid 3 (φ3)

18522.73

Apparent order (P)

4.45

Extrapolated values (φext21)

18125 21

Approximate relative error (ea )

0.46%

21

Extrapolated relative error (eext )

0.19%

Fine grid convergence index (CGIfine21)

0.24%

Table 3 Geometrical details of test rig Impeller Inlet Diameter

0.25 m

Impeller Exit Diameter

0.66 m

Impeller width

0.05 m

Inlet Blade angle

40°

Exit Blade angle

55°

Number of blades

24

Inlet duct Diameter

0.25 m

Volute width

0.10 m

Speed

1800 rpm

Table 4 Validation data

Flow coefficient (φ)

Total-Total isentropic Efficiency (η)

work done factor (Ψ)

Experimental

CFD

Experimental

CFD

0.14

52.40

58.16

0.48

0.48

0.18

53.40

56.09

0.42

0.41

0.20

49.70

52.48

0.37

0.37

0.20

49.30

51.95

0.36

0.36

Graphical Abstract

tongue

inlet duct

top wall Parallel wall side wall

Rectangular

0.5

C

pressure recovery coefficient

90°

0.5

0.4

C RE 5.0

CP RE 4.0

C RE 4.0

CP R 5.0

C R 5.0

CP R 4.0

C R 4.0

0.8

0.9

1.0

normalized mass flow rate (m/md)

Volute Performance

I

Normalized axial distance (Z/B2)

(a)

(N/m2)

(b)

I

Normalized radial distance (R/R2) Total pressure distribution

0.4

0.3

0.3 0.7

Different shape volute

CP RE 5.0

0.2 1.1

loss coefficient

separator wall

270°

0.6

CP

0.6

Highlights  An attempt made to explore energy efficient volute for industrial blowers  Parallel wall volutes assessed for aerodynamic performance against rectangular  Increased efficiency of 6% achieved with R 4.0 parallel wall volute  Parallel wall volutes can be an efficient and effective alternative