Design of Heat Transfer Equipment

Transcription

1 Design of Heat Transfer Equipment

2 Types of heat transfer equipment Type service Double pipe exchanger Heating and cooling Shell and tube exchanger All applications Plate heat exchanger Plate-fin exchanger Heating and cooling Spiral heat exchanger Air cooled Cooler and condensers Direct contact Cooling and quenching Agitated vessel Heating and cooling Fired heaters Heating 2

3 Some more terminology Exchanger: heat exchanged between two process streams Heaters and coolers: where one stream is plant service Vaporizer: if a process stream is vaporized Reboiler: a vaporizer associated with distillation column Evaporator: if concentrating a solution Fired exchanger: if heated by combustion gases 3 Unfired exchanger: not using combustion gases

4 BASIC THEORY General equation for heat transfer across a surface is Q = heat transferred per unit time, W U = the overall heat transfer coefficient, W/m 2o C A = heat-transfer area, m 2 T m = the mean temperature difference, o C Q UA T m 4

5 Geometry Resistances to heat transfer outside inside 5

6 TRANSFER COEFFICIENTS U o =overall coefficient on outside area of tube, W/m 2 o C h o =outside film coefficient, W/m 2 o C h i =inside film coefficient, W/m 2 o C h od =outside dirt coefficient, W/m 2 o C h id =inside dirt coefficient, W/m 2 o C k w =thermal conductivity of wall material, W/m o C 6

7 OVERALL HEAT TRANSFER COEFFICIENT The overall coefficient is reciprocal of the overall resistance to heat transfer, which is the sum of several individual resistances. Individual resistance is the reciprocal of individual HTC. d d ln o o d i do 1 do 1 U h h 2k d h d h o o od w i id i i 7

8 COMMENTS Magnitude of h s depends on: nature of the process (conduction, convection, radiation, condensation, etc.) Physical properties (density, heat capacity, viscosity, thermal conductivity) Fluid flow rates Physical arrangement of exchanger 8

9 Typical design procedure 1-Define the duty: Q, F s, T s 2- Physical properties required: density, viscosity, thermal conductivity 5- t m 6-Calculate area required Q UA T m 4-trial value of OHTC (U) 7-Exchanger layout 3- Type of exchanger 9-Calculate the OHTC Uc U d o d ln 9 1 U o 1 h o 1 h 10-Calculate p od o d 2k w i d d o i 1 h id d d o i 1 h i 8-h i and h o 11-Optimise: repeat steps 4-10

10 To select a trial value of U Select a trial Value of U for given fluids. 10

11 Fouling or dirt factor What? Deposit of nonmetallic material on heat transfer surface is fouling Consequences Heat transfer resistance is increased which require over design of exchanger 11

12 Shell & Tube Exchangers Advantages Large surface area per unit volume Uses well established fabrication techniques Can be constructed from a wide range of materials Easy cleaning 12

13 Selecting TEMA Type Heat Exchangers Tubular Exchange Manufacturers Association The general descriptions of the three major TEMA classes are: TEMA C - General Service TEMA B - Chemical Service TEMA R - Refinery Service TEMA R is the most restrictive and TEMA C is the least stringent. TEMA B and TEMA R are very similar in scope. TEMA R requires a greater minimum thickness for some components. 13

14 Straight Tube, Fixed Tube-sheet, Type BEM, AEM, NEN, Etc. This TEMA type is the simplest design and is constructed without packed or gasketed joints on the shell side. The tube-sheet is welded to the shell and the heads are bolted to the tube-sheet. On the NEN heat exchanger, the shell and the head is welded to the tube-sheet. Typically, a cover plate design is provided to facilitate tube cleaning. This TEMA category, especially the NEN, it is the lowest cost TEMA design per square foot of heat transfer surface. 14

15 Advantages Less costly than removable bundle designs Provides maximum amount of surface for a given shell and tube diameter Provides for single and multiple tube passes to assure proper velocity May be interchangeable with other manufacturers of the same TEMA type Limitations Shell side can be cleaned only by chemical methods No provision to allow for differential thermal expansion, must use an expansion joint Applications Oil Coolers, Liquid to Liquid, Vapor condensers, reboilers, gas coolers Generally, more viscous and warmer fluids flow through the shell Corrosive or high fouling fluids should flow inside the tubes 15

16 Removable Bundle, Externally Sealed Floating Tube-sheet, Type OP, AEW, BEW This design allows for the removal, inspection and cleaning of the shell circuit and shell interior. Special floating tube-sheet prevents intermixing of fluids. In most cases, straight tube design is more economical than U-tube designs. Advantages Floating tube-sheet allows for differential thermal expansion between the Shell and the tube bundle. Shell circuit can be inspected and steam or mechanically cleaned The tube bundle can be repaired or replaced without disturbing shell pipe Less costly than TEMA type BEP or BES which has internal floating head Maximum surface for a given shell diameter for removable bundle design Tubes can be cleaned in AEW models without removing 16

17 Limitations Fluids in both the shell and tube circuits must be nonvolatile, non-toxic Tube side passes limited to single or two pass design All tubes are attached to two tube-sheets. Tubes cannot expand independently so that large thermal shock applications should be avoided Packing materials produce limits on design pressure and temperature Applications Intercoolers and after-coolers, air inside the tubes Coolers with water inside the tubes Jacket water coolers or other high differential temperature duty Place hot side fluid through the shell with entry nearest the front end 17

18 Removable Bundle, Outside Packed Head, Type BEP, AEP, Etc This design allows for the easy removal, inspection and cleaning of the shell circuit and shell interior without removing the floating head cover. Special floating tube-sheet prevents intermixing of fluids. In most cases, straight tube removable design is more costly than U-tube designs. 18

19 Advantages Floating tube-sheet allows for differential thermal expansion between the shell and the tube bundle. Shell circuit can be inspected and steam cleaned. If the tube bundle has a square tube pitch, tubes can be mechanically cleaned by passing a brush between rows of tubes. The tube bundle can be repaired or replaced without disturbing shell piping On AEP design, tubes can be serviced without disturbing tubeside piping Less costly thantema type BES or BET designs Only shell fluids are exposed to packing. Toxic or volatile fluids can be cooled in the tubeside circuit Provides large bundle entrance area, reducing the need for entrance domes for proper fluid distribution 19

20 Limitations Shell fluids limited to non volatile, non toxic materials Packing limits shell side design temperature and pressure All tubes are attached to two tube-sheets. Tubes cannot expand independently so that large thermal shock applications should be avoided Less surface per given shell and tube diameter than AEW or BEW Applications Flammable or toxic liquids in the tube circuit Good for high fouling liquids in the tube circuit 20

21 Removable Bundle, Internal Split Ring Floating Head, Type AES, BES, Etc. - Ideal for applications requiring frequent tube bundle removal for inspection and cleaning. Uses straight-tube design suitable for large differential temperatures between the shell and tube fluids. More forgiving to thermal shock than AEW or BEW designs. Suitable for cooling volatile or toxic fluids. 21

22 Advantages Floating head design allows for differential thermal expansion between the shell and the tube bundle. Shell circuit can be inspected and steam cleaned. If it has a square tube layout, tubes can be mechanically cleaned Higher surface per given shell and tube diameter than pull-through designs such as AET, BET, etc. Provides multi-pass tube circuit arrangement. Limitations Shell cover, split ring and floating head cover must be removed to remove the tube bundle, results in higher maintenance cost than pull-through More costly per square foot of surface than fixed tube sheet or U-tube designs Applications Chemical processing applications for toxic fluids Special intercoolers and after-coolers General industrial applications 22

23 Removable Bundle, Pull-Through Floating Head, Type AET, BET, etc. Ideal for applications requiring frequent tube bundle removal for inspection and cleaning as the floating head is bolted directly to the floating tube-sheet. This prevents having to remove the floating head in order to pull the tube bundle. Advantages Floating head design allows for differential thermal expansion between the shell and the tube bundle. Shell circuit can be inspected and steam or mechanically cleaned Provides large bundle entrance area for proper fluid distribution Provides multi-pass tube circuit arrangement. Suitable for toxic or volatile fluid cooling 23

24 Limitations For a given set of conditions, this TEMA style is the most expensive design Less surface per given shell and tube diameter than other removable designs Applications Chemical processing applications for toxic fluids Hydrocarbon fluid condensers General industrial applications requiring frequent cleaning 24

25 Removable Bundle, U-Tube, Type BEU, AEU, Etc. Especially suitable for severe performance requirements with maximum thermal expansion capability. Because each tube can expand and contract independently, this design is suitable for larger thermal shock applications. While the AEM and AEW are the least expensive, U-tube bundles are an economical TEMA design. 25

26 Advantages U-tube design allows for differential thermal expansion between the shell and the tube bundle as well as for individual tubes. Shell circuit can be inspected and steam or mechanically cleaned Less costly than floating head or packed floating head designs Provides multi-pass tube circuit arrangement. Capable of withstanding thermal shock applications. Bundle can be removed from one end for cleaning or replacement Limitations Because of u-bend, tubes can be cleaned only by chemical means Because of U-tube nesting, individual tubes are difficult to replace No single tube pass or true countercurrent flow is possible Tube wall thickness at the U-bend is thinner than at straight portion of tubes Draining of tube circuit is difficult when mounted with the vertical position With the head side up. Applications Oil, chemical and water heating applications Excellent in steam to liquid applications 26

27 TUBES DIMENSIONS Range: 16 mm to 50 mm Common size: 16 mm-25 mm OD and wall thickness shall be specified or nominal size Lengths: 1.83 m, 2.44 m,3.66 m,4.88 m,6.10 m,7.32 m 19 mm OD is good starting value 27

28 TUBE ARRANGEMENTS A. Equilateral triangular P 1 B. Square C. Rotated square P 1 A & C give higher HTC A & C give higher P.D. C is for fouling liquids(easy mechanical cleaning) Tube pitch=1.25od 28

29 TUBE SIDE PASSES One tube pass Two tube pass Three tube passes 29

30 TUBE SHEET LAYOUT Tube bundle diameter depends on: Number of tubes Number of tube passes Estimate of Bundle Diameter can be obtained from Eqn. N t =number of tubes D b =bundle diameter, mm d o =outside tube diameter, mm 30

31 D b d o N K t 1 1/ n 1 31

32 Baffles Purpose: Direct flow across the tubes Increase heat transfer 32

33 Temperature Mean Temperature Difference T 1 t 2 T 2 T 1 t 2 shell T 2 t 1 t 1 tubes Heat transferred t 2 T 2 Heat transferred T 2 t 2 t 1 33 t 1 T T 1 1 T 1 t2 T 2 t1 T 1 t1 T 2 t2 T lm T ln T 1 2 t t 2 1 T lm T ln T Counter-current Co-current 1 2 t t 1 2

34 LMTD t F t F t R m t lm f ( R, S) T T t t, S t t T t

35 35

36 R = (m s X C P,f,s / m t X C P,f,t ) S is a measure of the temperature efficiency of the exchanger. 36

37 The following assumptions are made in the derivation of the temperature correction factor F t, in addition to those made for the calculation of the log mean temperature difference: 1. Equal heat transfer areas in each pass. 2. A constant overall heat-transfer coefficient in each pass. 3. The temperature of the shell-side fluid in any pass is constant across any cross-section. 4. There is no leakage of fluid between shell passes 37

38 Design Considerations: STE Fluid Allocation: Shell Or Tubes Where no phase change occurs, Corrosion Fouling Fluid Temperatures Operating pressure Pressure drop Viscosity Stream flow rates 38

39 Corrosion The more corrosive fluid should be allocated to the tube-side. This will reduce the cost of expensive alloy or clad components. 39

40 Fouling The fluid that has the greatest tendency to foul the heattransfer surfaces should be placed in the tubes. This will give better control over the design fluid velocity, The higher allowable velocity in the tubes will reduce fouling. Also, the tubes will be easier to clean. 40

41 Fluid Temperatures If the temperatures are high enough to require the use of special alloys placing the higher temperature fluid in the tubes will reduce the overall cost. At moderate temperatures, placing the hotter fluid in the tubes will reduce the shell surface temperatures, and hence the need for lagging to reduce heat loss, or for safety reasons, 41

42 Operating Pressure The higher pressure stream should be allocated to the tube-side. High-pressure tubes will be cheaper than a high-pressure shell. 42

43 Pressure Drop For the same pressure drop, higher heat-transfer coefficients will be obtained on the tube-side than the shell-side, and fluid with the lowest allowable pressure drop should be allocated to the tube-side. 43

44 Viscosity Generally, a higher heat-transfer coefficient will be obtained by allocating the more viscous material to the shell-side, providing the flow is turbulent. The critical Reynolds number for turbulent flow in the shell is in the region of 200. If turbulent flow cannot be achieved in the shell it is better to place the fluid in the tubes, as the tube-side heat-transfer coefficient can be predicted with more certainty. 44

45 Stream Flow Rates Allocating the fluids with the lowest flow-rate to the shell-side will normally give the most economical design. 45

46 Design Considerations: STE Fluid Velocities Liquids Tube side Process fluids 1-2 m/s 4 m/s maximum Water m/s Shell side m/s Vapors Vacuum m/s Atmospheric pressure m/s 46

47 PRESSURE DROP Liquids Viscosity <1 mn s/m 2 p = 35 kn/m 2 Viscosity is 1 to 10 mn s/m 2 p = kn/m 2 Gas and vapours High vacuum p = kn/m 2 Medium vacuum p = 0.1 x absolute pressure 1 to 2 bar p = 0.5 x system gauge pressure Above 10 bar p = 0.1 x system gauge pressure 47

48 Tube-side Heat-transfer Coefficient And Pressure Drop (Single Phase) Heat transfer Turbulent flow Heat-transfer data for turbulent flow inside conduits of uniform cross-section are correlated by an equation of the form: 48

49 where Nu = Nusselt number = (h i d e /k f ), Re = Reynolds number = (ρu t de/µ) = (G t d e /µ), Pr = Prandtl number = (Cpµ/k f ) h i = inside coefficient, W/m 2 C, d e = equivalent (or hydraulic mean) diameter, m = (4 x cross-sectional area for flow/ wetted perimeter ) = d i for tubes, 49

50 u t = fluid velocity, m/s, k f = fluid thermal conductivity, W/m C, G t = mass velocity, mass flow per unit area, kg/m 2 s, µ = fluid viscosity at the bulk fluid temperature, Ns/m 2, µ w = fluid viscosity at the wall, C p = fluid specific heat, heat capacity, J/kg C. 50

51 Now, a = 0.8. b = 0.3 for cooling = 0.4 for heating. c = 0.14 for flow in tubes. 51

52 A general equation that can be used for exchanger design is: where C = for gases, = for non-viscous liquids, = for viscous liquids. 52

53 Butterworth (1977) gives the following equation, Where St = Stanton number = (Nu/Re Pr) = (hi/ρµ t C P ) And E = exp( (ln Pr) 2 ) This equation is applicable at Reynolds numbers greater than 10,000 53

54 Hydraulic Mean Diameter For turbulent flow in a duct of non-circular cross-section, the hydraulic mean diameter may be used in place of the pipe diameter and the formulae for circular pipes can then be applied without introducing a large error. This method of approach is entirely empirical. The hydraulic mean diameter D H is defined as four times the hydraulic mean radius r H. Hydraulic mean radius is defined as the flow cross-sectional area divided by the wetted perimeter: 54

55 Some examples are given. For circular pipe: D H = 4(π/4)D 2 / (π D) = D For an annulus of outer diameter D o and inner diameter D i : D H = 4 ( (π D o2 /4) - (π D i2 /4) ) / ( π(d o + D i ) ) = (D o2 - D i2 ) / (D o + D i ) = D o - D i For a duct of rectangular cross-section D a by D b : D H = 4 D a D b / ( 2(D a + D b ) = 2D a D b / (D a + D b ) For a duct of square cross-section of size D a : D H = 4 D a2 / (4D a ) = D a For laminar flow this method is not applicable, and exact expressions relating the pressure drop to the velocity can be obtained for ducts of certain shapes only. 55

56 Laminar flow Below a Reynolds number of about 2000 the flow in pipes will be laminar. Providing the natural convection effects are small, which will normally be so in forced convection, to estimate the film heattransfer coefficient given equation will be used: where L is the length of the tube in metres. If the Nusselt number given by above equation is less than 3.5, it should be taken as

57 Heat-transfer factor, j h : It is often convenient to correlate heat-transfer data in terms of a heat transfer j h factor. The heat-transfer factor is defined by: The use of the j h factor enables data for laminar and turbulent flow to be represented on the same graph. 57

58 58

59 Equation can be rearranged to a more convenient form: Kern (1950)define the heat transfer factor as: The relationship between j h and J H is given by: 60

60 Tube-side pressure drop There are two major sources of pressure loss on the tube-side of a shell and tube exchanger: The friction loss in the tubes and The losses due to the sudden contraction and expansion and minor source of pressure loss : flow reversals that the fluid experiences in flow through the tube arrangement. The tube friction loss can be calculated using the familiar equations for pressure-drop loss in pipes. (see fluid mechanics). 61

61 The basic equation for isothermal flow in pipes (constant temperature) is: where j f is the dimensionless friction factor and L' is the effective pipe length. 62

62 The flow in a heat exchanger will clearly not be isothermal, and this is allowed for by including an empirical correction factor to account for the change in physical properties with temperature. Normally only the change in viscosity is considered: Values of j f for heat exchanger tubes can be obtained from Figure. 63

63 64

64 The pressure losses due to contraction at the tube inlets, expansion at the exits, and flow reversal in the headers, can be a significant part of the total tube-side pressure drop. There is no entirely satisfactory method for estimating these losses. Kern (1950) suggests adding four velocity heads per pass. Frank (1978) considers this to be too high, and recommends 2.5 velocity heads. Butterworth (1978) suggests 1.8. Lord et al. (1970) take the loss per pass as equivalent to a length of tube equal to 300 tube diameters for straight tubes, and 200 for U- tubes; whereas Evans (1980) appears to add only 67 tube diameters per pass. 65

65 The loss in terms of velocity heads can be estimated by counting the number of flow contractions, expansions and reversals, and using the factors for pipe fittings to estimate the number of velocity heads lost. For two tube passes, there will be two contractions, two expansions and one flow reversal. The head loss for each of these effects is: contraction 0.5, expansion 1.0, 180 bend 1.5; so for two passes the maximum loss will be 2 x x = 4.5 velocity heads = 2.25 per pass 66

66 From this, it appears that Frank's recommended value of 2.5 velocity heads per pass is the most realistic value to use. 67

67 Viscosity Correction Factor The viscosity correction factor will normally only be significant for viscous liquids. 68

68 Coefficients For Water The equation below has been adapted from data given by Eagle and Ferguson (1930): where h i = inside coefficient, for water, W/m 2 C, t = water temperature, C, u t = water velocity, m/s, d i = tube inside diameter, mm. 69

69 Shell-Side Heat-Transfer And Pressure Drop (Single Phase) 70

70 71

71 DESIGN METHODS Kern Method Bell or Bell-Delware Method 72

72 KERN'S METHOD The shell equivalent diameter is calculated using the flow area between the tubes taken in the axial direction (parallel to the tubes) and the wetted perimeter of the tubes; 73

73 Procedure 1. Calculate the area for cross-flow As for the hypothetical row of tubes at the shell equator, given by: 74

74 2. Calculate the shell-side mass velocity G s and the linear velocity u s : where W s = fluid flow-rate on the shell-side, kg/s, ρ = shell-side fluid density, kg/m 3. 75

75 3. Calculate the shell-side equivalent diameter (hydraulic diameter), D D D e e e 4 free flow area wetted perimeter 4 P 4 P 2 T 2 T d d o 2 o / 4 Where D e = equivalent diameter in m 3 / 4 d d / 2 o 2 o square / 8 triangular 76

76 4. Calculate the shell-side Reynolds number, given by: 77

77 5. For the calculated Reynolds number, read the value of j h, from graph for the selected baffle cut and tube arrangement, and calculate the shell-side heat transfer coefficient h s from: 78

78 6. For the calculated shell-side Reynolds number, read the friction factor from graph and calculate the shell-side pressure drop from: where L = tube length, I B = baffle spacing. The term (L/l B ) is the number of times the flow crosses the tube bundle = (N b + 1), where N b, is the number of baffles. 79

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81 Heat Exchanger Shell Side Design Bell Method

82 SHELL-SIDE HEAT-TRANSFER AND PRESSURE DROP (SINGLE PHASE) Flow pattern

83 Bell s method In Bell s method the heat-transfer coefficient and pressure drop are estimated from correlations for flow over ideal tubebanks, and the effects of leakage, bypassing and flow in the window zone are allowed for by applying correction factors.

84 Heat-transfer coefficient The total correction will vary from 0.6 for a poorly designed exchanger with large clearances to 0.9 for a well-designed exchanger.

85 h oc, ideal cross-flow coefficient

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87 F n, tube row correction factor

88 F w, window correction factor

89 F b, bypass correction factor With sealing strips

90 Where no sealing strips are used, F b can be obtained from Figure

91 F L, Leakage correction factor

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93 Shell and bundle geometry

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99 Pressure drop Cross-flow zones The pressure drop in the cross-flow zones between the baffle tips is calculated from correlations for ideal tube banks, and corrected for leakage and bypassing.

100 Pressure Drop ideal tube bank pressure drop The number of tube rows has little effect on the friction factor and is ignored.

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102 F b, bypass correction factor for pressure drop Bypassing will affect the pressure drop only in the cross-flow zones. The correction factor is calculated from the equation used to calculate the bypass correction factor for heat transfer,

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104 F L, leakage factor for pressure drop

105 Window-zone pressure drop

106 End zone pressure drop

107 Total shell-side pressure drop

108 THANK YOU