numerical analysis of spiral wound heat exchanger based on fluid … · 2018-01-07 · 4th...

4th International Conference On Building Energy, Environment

Numerical analysis of spiral wound heat exchanger based on fluid-structure interaction

Simin Wang1, Guanping Jian1, Lijuan Sun1, Jian Wen2

1 School of chemical engineering and technology Xi'an Jiaotong University, Xi'anShanxi, 710049 China

2 School of energy and power engineering Xi'an Jiaotong University, Xi'an, 710049 China

SUMMARY Based on the fluid-structure coupling method, the effects of configuration parameters on the flow characteristic, the heat transfer performance and the stress distribution on the tube bundle of spiral wound heat exchanger (SWHE) were numerically studied. The results show that the shell-side flow patterns change from cross flow into oblique flow and the overall heat transfer coefficient increases firstly and then decreases with the increase of the wingding angle. The maximum stress intensity decreases firstly with the increase of the inlet flow rate and then decreases. The thermal stress takes a leading position in the tube bundle stress distribution attributed to the higher heat transfer temperature difference. With the inlet flow rate increasing, the heat transfer is enhanced and the temperature difference decreases which result in the decrease of the proportion of thermal stress and the increase of the impact of the primary stress.

INTRODUCTION Spiral wound heat exchangers (SWHE) are widely used in liquefied natural gas plants, air separation plants, petrochemical enterprises, nuclear power stations due to their advantages of high-pressure endurance, highly compact structure and good thermal compensation performance [Zhang 2008, Weikl et al. 2014]. The structure of SWEH, which comprises multiples layers of tubes spirally wound around a cylinder, is quite complicated and the winding direction of the tubes vary for each tube layer to achieve a compact design with comparatively long tube lengths. Winding the tubes in the shell side of the SWEH is a well know effective method to enhance the heat transfer performance and a lot of researchers have investigated the heat transfer and flow characteristics in the shell side of SWEH. Nessrass et al. (2004) constructed a test plant to measure the local heat transfer coefficients and frictional pressure drop and studied the liquid falling film flow in the shell side of spiral wound LNG heat exchanger. Based on the obtained experimental data, the heat transfer coefficients were compared with the data from Gnielinski (1998) and MaAdams et al. (1937) Furthermore, a modified method from Barbe et al. (1972)for the frictional pressure drop was recommended as well. Charles J et al.(1969), Ghorbani et al.(2010) and Jamshidi et al.(2013) experimentally analyzed the effects of the geometrical parameters on heat transfer performance. And they found that the structural parameters had significant effects on flow and heat transfer. Along with the development of CFD and computer technology, the numerical simulations about the study of heat exchanger get really popular. Ferng et al. (2012) studied the effects of

different Dean number and pitch size on the thermal-hydraulic characteristics in a helically coil-tube heat exchanger based on the numerical simulation and reasonably captured the complicated phenomena occurred in a helically coiled-tube heat exchanger. Wang et al. (2016) proposed a simulation tool of floating LNG SWHE and analyzed the impact of rolling amplitude on the heat transfer performance. Haskins et al. (2016) numerically analyzed the friction pressure drop of isothermal flows on the shell side of annular heat exchangers with helically coiled concentric tubes and the developed correlation agreed with the CFD results was presented. Over the last years, little attention has been concentrated on the study both of thermal-hydraulic performance and stress distribution in spiral wound heat exchanger on account of fluid -structural analysis. In this paper, the parameter drive design would be applied to investigate the flow and thermal performances of different geometrical structures and the effects of geometrical parameters (the winding angle, the tube pitch, the external the diameter of the tube and wall thickness of the tube) on characteristics of thermal and stress distribution of the tube bundle.

METHODS

Figure 1. Structure of spiral wound heat exchanger

The schematic diagram of a spiral wound heat exchanger is shown in Figure 1. Geometrical parameters (the winding angle θ, the tube pitch l, the external diameter of the tube Dt and wall thickness of the tube T) would be set as drive parameters to achieve parameter drive design and the different structures of SWHX would be obtained by changing the drive parameters. To capture the stress distribution of the tube bundle and improve the accuracy of the simulation, four spacing bar would be circumferential uniformly distributed in the physical model (Fig. 1). The ranges of the geometrical parameters used in this study are listed in Table 1. The diameter of the first winding layer is 241mm and the effective length of the heat exchange of tube bundle is 320mm.

ISBN: 978-0-646-98213-7 COBEE2018-Paper340 page 948


Table 1. The variation of geometrical parameters

Geometrical parameters Variation Range

Winding Angle θ / ° 10-22

Wall Thickness T / mm 0.8-2.5

Tube Pitch l / mm 5-12

External Diameter of Tubes Dt / mm 9-20

Due to the complex geometrical structure of SWHE, the computational domain is meshed with unstructured grids and partial refinement is made near the wall of the calculation domain. Mesh adaption and grid independence tests are conducted to ensure the accuracy of the numerical results. In view of the different size of various geometries, the grid numbers for spiral-wound heat exchangers were 5,000,000-18,000,000 in this paper. The working fluid in this study is water and its physical properties are considered to be constant. The inlet and outlet are set as velocity inlet and pressure outlet. The inlet velocity of the tube side is 1 m·s-1 and the temperature of the water is 86℃. The inlet velocity of the shell side is varied from 0.5 m·s-

1 to 2.5 m·s-1 and the temperature of the inlet fluid is 37℃. The first order high resolution scheme is used to discretize the convection terms. The SIMPLEC method is adopted for the coupling of pressure and velocity. The convergence criterion is that the normalized residuals are less than 1 × 10-6 for the flow equations and 1 × 10-8 for the energy equation. After the calculation of the flow field reaching the convergence criterion, both the pressure and temperature load of the tube bundle would be mapped to the solid domain. According to the manufacturing technology, both ends of the tube bundle are set as fix support boundary condition. The material of the tubes and spacing bar is structural steel. The Nu and △ P obtained by numerical simulation are compared with the experimental data in reference [4], as show in Figure 2. The average discrepancy of Nusselt number is 5.67% and the discrepancy of pressure drop per unit length is 7.90%, which can be accepted in the engineering applications.

30000 40000 50000 60000

75

150

225

300

Re

Nu

500

1000

1500

2000

2500

Nu (Simulation) Nu (Experiment[4]) ∆Pm (Simulation) ∆Pm (Experiment[4])

∆Pm (P

a/m

)

Figure 2. Comparison of the experimental data and the numerical simulation

RESULTS EFFECTS OF THE GEOMETRICAL PARAMETERS ON PRESSURE DROP Figure 3 shows the variation of the shell-side pressure drop with winding angle under different inlet velocity. It can be noted that the shell-side pressure drop decreases with the increase

of the winding angle under the specified inlet velocity. The shell-side flow patterns change with the winding angle. The shell-side fluid is cross flow along the tubes at a very small winding angle. The angle between the flow direction of the fluid and the axis of the tube is near to 90°. When the winding angle is large, the cross flow turns into oblique flow and the angle between flow direction and axis of the tube increases, which results in the decrease of flow loss.

10 12 14 16 18 20 220

5000

10000

15000

20000

25000

30000 T = 1.65mml = 8.5mm

Dt = 14.5mm

Pres

sure

dro

p (P

a)

Winding angle (°)

0.5 m/s 1 m/s 1.5 m/s

2 m/s 2.5 m/s

Figure 3. The shell-side pressure drop versus winding angle under different inlet velocity

Figure 4 shows the variation of the shell-side pressure drops with the shell inlet velocity under different external tube diameters. When the inlet velocity is less than 1.25 m·s-1, the change of the external diameter has a smaller impact on the shell-side pressure. However, the pressure drop increases with the increase of the external diameter when the inlet velocity is larger than 1.5 m·s-1. The flow field disturbance intensifies when the velocity in the shell side is higher, which means it would result in larger flow losses. Furthermore, the equivalent flow area of the shell side decreases with the increase of the external diameter and it would increase the shell-side pressure drop as well

0.5 1.0 1.5 2.0 2.50

5000

10000

15000

20000

25000

30000

35000

Pres

sure

Dro

p (P

a)

Vin( m·s-1)

Dt= 9 mm Dt= 12.7 mm Dt= 16.3 mm Dt= 20 mm

T =16.5mml = 8.5mmθ = 16°

Figure 4. The shell-side pressure drop versus inlet velocity under different external tube diameter

EFFECTS OF THE GEOMETRICAL PARAMETERS ON THE HEAT TRANSFER Figure 5 presents the overall heat transfer coefficient versus winding angle under different inlet velocity. The results illustrate that the overall heat transfer coefficient increases firstly and then decreases as winding angle increase. Under the specified winding angle, there is a maximum value of the overall heat transfer coefficient. It can be explained by the flow pattern transition in the shell-side when the winding angle is changed. Under small winding angle, the flow pattern in the shell-side is considered as cross flow. But the downstream tubes may locate in the wake region of upstream adjacent tubes. With the increase of the winding angle, the diversion effect of the winding tubes would facilitate the fluid to scour



the wake region and the local heat transfer would be improved. The diversion effect becomes more obvious with the further increase of the winding angle and dominates the whole flow field. When the flow pattern turns into oblique flow, the overall heat transfer would decrease with the increase of winding angle. As a result, when the winding angle increases, the overall heat transfer coefficient increases firstly and then decreases.

10 12 14 16 18 20 22

2700

2800

2900

3000

K(W

·m-2·K

-1)

θ ( ° )

Vin=1.5 m·s-1

Vin=1.7 m·s-1

Vin=1.9 m·s-1

Vin=2.2 m·s-1

T=1.65mml=8.5mmDt=14.5mm

Figure 5. The overall heat transfer coefficient versus winding angle under different inlet velocity

8 10 12 14 16 18 202400

2600

2800

3000

K(W

·m-2·K

-1)

Dt (mm)

Vin=1.5 m·s Vin=1.7 m·s Vin=1.9 m·s Vin=2.2 m·s

T=1.65mmθ=18°l=8.5mm

Figure 6. The overall heat transfer coefficient versus external diameter under different inlet velocity

The overall heat transfer coefficient versus external diameter under different inlet velocity is given on Figure 6. As the figure shows, the overall heat transfer coefficient increases to the maximum and then decreases with the external diameter. The equivalent flow area becomes smaller when the external diameter increases and it will facilitate the heat transfer. But with the further increase of the external diameter, the proportion of areas of the downstream tubes which locate in the wake region will increase and it will result in the decrease of the overall heat transfer. STRESS ANALYSIS Stresses in the heat exchanger are caused by the temperature difference and the pressure load from fluid. It is necessary to determine what type of load dominating the stress of the tube bundle. The specific structure of the SWHE would be chosen to investigate the characteristic of the stress distribution under the different types of loads (winding angle=16°, external diameter=14.5mm, wall thickness=1.62mm, tube pitch=14.5mm). Three different types of load would be applied to the tube bundle, which are only pressure load, only temperature difference loads, both the pressure load and the temperature difference load, respectively. The results are shown in Table 2.

Table 2 The maximum of stress intensity under the different types of load and shell-side inlet velocity (Winding angle=16°, External diameter=14.5mm, Wall thickness=1.62mm, Tube pitch=14.5mm)

Type of load Velocity

m·s-1 Maximum of stress Intensity

MPa

Pressure

0.5 1.5005 1.5 13.561 2.5 37.187

Temperature difference

0.5 236.02

1.5 176.12

2.5 159.32

Pressure& Temperature

difference

0.5 237.52 1.5 189.68 2.5 196.5

Table 3 The force, temperature difference and K under the different shell-side inlet velocity

Velocity m·s-1

Temperature Difference

K

K W·m-

2·K-1

Force in Y direction

N 0.5 42.2 1872.9 9.75 1.5 37.1 2762.1 95.6 2.5 36.2 3053.8 262.4

The force in Y direction (axis direction of the heat exchanger) and the overall heat transfer coefficient increase with the increase of the inlet velocity and they are equal to each other under the different types of load (in Table 3). The temperature difference decreases when the inlet velocity increases. It means that the pressure load increases with the increase of the inlet velocity while the changes of the temperature difference load are opposite. As illustrated in Table 2, the maximum of the stress intensity of the pressure load under specific inlet velocity is far less than that of the temperature difference load. It means the thermal stress dominates the stress distribution of the tube bundle but the proportion of the proportion of primary stress would increase when the inlet velocity increases. In this paper, both pressure load and temperature difference load would map on the tube bundle to ensure the accuracy of the numerical result.

0.5 1.0 1.5 2.0 2.5

150

160

170

180

190

200

210

220

S (M

Pa)

Vin (m·s-1)

T=0.8mm T=1.4mm T=1.9mm T=2.5mm

θ=16°Dt=14.5mml=8.5mm

Figure 7. Maximum stress intensity of tube bundle versus inlet velocity under the different wall thickness

Figure 7 and Figure 8 show the maximum stress intensity with inlet velocity under the different wall thickness and the different external diameter. In specific structure of the SWHE, the maximum stress intensity decreases firstly with inlet velocity then increases. The temperature difference of heat



exchange at low inlet velocity is smaller than that at the high inlet velocity, while the force of the axis direction increases with the increase of the inlet velocity. It means that the thermal stress takes a leading position in the smaller inlet velocity but the proportion of the primary stress increases with the increase of inlet velocity. This mechanism can interpret the phenomenon that the maximum stress intensity decreases firstly with the increase of the inlet flow rate and then decreases.

0.5 1.0 1.5 2.0 2.5

160

170

180

190

200

210

S (M

Pa)

Vin (m·s-1)

Dt=11.75mm Dt=14.5mm Dt=17.25mm Dt=20 mm

T=1.65mmθ=16°l=8.5mm

Figure 8. Maximum stress intensity of tube bundle versus inlet velocity under the different external diameter

CONCLUSIONS Based on the fluid-structure analysis, the effects of geometrical parameters on thermal hydraulic performance and stress distribution in spiral wound heat exchanger were numerically investigated. With the increase of the winding angle, the diversion effect of the winding tubes becomes more obvious and the flow pattern in the shell side turns from the cross flow into the oblique flow. The heat transfer coefficient increases firstly with the increase of winding angle then decreases, while the pressure drop decreases with the increase of the winding angle. With the increase of the external diameter, the shell-side equivalent flow area and the proportion of the downstream-tube areas in the wake region vary. Then the overall heat transfer coefficient increases to the maximum and then decreases with the external diameter. The structural optimization to achieve both high heat transfer comprehensive performance and stability of operation was also conducted as well. The thermal stress dominates the stress distribution of the tube bundle. With the increase of inlet flow rate, the heat transfer is enhanced and the temperature difference decreases, which results in the decrease of the proportion of thermal stress and strengthens the impact of the primary stress.

ACKNOWLEDGEMENT This work is supported by the National Natural Science Foundation of China (No. 51676146).

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numerical analysis of spiral wound heat exchanger based on fluid … · 2018-01-07 · 4th...

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