Low-GWP refrigerant heat exchange optimization design method

Abstract

The invention discloses a low-GWP refrigerant heat exchange optimization design method, and relates to the technical field of fire safety, an ecological design method of a heat regeneration heat exchange tube based on a refrigerant GWP value comprises the following steps: firstly, selecting a refrigerant type, determining a refrigeration system preset standard meeting ecological environment requirements, and then selecting proper heat regeneration heat exchange tube tubing and specification, so as to obtain a heat regeneration heat exchange tube material; according to the method, the internal thread parameter value is considered while the tube pass side regenerative heat exchange coefficient is calculated, so that the calculation of the heat transfer area of the regenerative heat exchange tube is more accurate; the effects of reducing the pipe length, reducing the pressure drop of a refrigerant on the shell pass side, reducing energy waste and saving manufacturing cost are achieved while the heat exchange amount is increased and the heat exchange efficiency is improved, and heat exchange equipment which is more energy-saving, excellent, controllable, ecological and higher in heat exchange performance is provided for a user so as to meet the higher requirement of the modern society for a regenerative heat exchange device.

Description

Technical Field

[0001] The present invention relates to the field of heat recovery and heat exchange technology, and in particular to a low GWP refrigerant heat exchange optimization design method. Background Art

[0002] With the development of economy, heat recovery heat exchange tubes play an increasingly important role in the field of refrigeration and are more and more widely used. Therefore, the design requirements for heat recovery heat exchange tubes are becoming more and more stringent. The existing heat recovery heat exchange efficiency inside and outside the heat recovery heat exchange tube is not high. If the heat exchange efficiency is improved only by changing the structure, the refrigerant flow resistance inside the heat recovery heat exchange tube will increase. Therefore, how to optimize the structure of the heat recovery heat exchange tube to improve heat exchange while minimizing the refrigerant flow resistance and reducing energy consumption is also a difficulty in the current heat recovery heat exchange tube design field.

[0003] In order to achieve the lightweight of the heat recovery heat exchange tube, the invention patent application document CN115235283A directly replaces the copper material of the heat recovery heat exchange tube with aluminum alloy material, thereby achieving the purpose of lightweighting the heat recovery heat exchange tube. The first disadvantage is that it ignores the problem of poor performance of the internal threaded heat exchange tube caused by the change of heat exchange material. The second disadvantage is that it wants to enhance the heat exchange by optimizing the tooth height and helix angle parameters of the internal thread in the heat recovery heat exchange tube, but fails to provide a quantitative data source to support the optimization process, and its design accuracy is not high. Summary of the invention

[0004] In order to overcome the problem in the prior art that changing the material of the heat exchange tube leads to poor heat exchange performance of the internally threaded heat exchange tube and low design accuracy, the present invention proposes a low GWP refrigerant heat exchange optimization design method and device.

[0005] To achieve the above object, the present invention adopts the following technical solutions, including:

[0006] A low GWP refrigerant heat exchange optimization design method, comprising:

[0007] S1: Select the type of refrigerant and determine the GWP of the refrigerant; GWP is the global warming potential;

[0008] S2: Calculate the carbon footprint energy efficiency value G based on the GWP of the refrigerant;

[0009] S3: Determine whether the carbon footprint value and the energy efficiency value of the refrigeration system are compatible, that is, whether the carbon footprint energy efficiency value meets the preset standard. If yes, proceed to the next step; otherwise, adjust the subcooling degree of the heat recovery heat exchange tube so that the carbon footprint energy efficiency value meets the preset standard, and then proceed to the next step;

[0010] S4: Based on the basic data of the condenser side and the evaporator side, estimate the heat transfer area and select appropriate pipe materials and specifications for the evaporator side and the condenser side;

[0011] S5: Continue to optimize the internal thread design of the heat recovery heat exchange tube to determine whether the optimized tube length and the optimized pressure drop on both sides meet the standard conditions. If yes, obtain the final design result of the heat recovery heat exchange tube. If not, re-evaluate the internal thread parameters of the heat recovery heat exchange tube and continue to optimize the design.

[0012] A heat recovery heat exchange tube device comprises: a heat recovery heat exchange tube, wherein a sleeve is arranged inside the heat recovery heat exchange tube, the sleeve adopts an internal thread structure, and the type of refrigerant and the internal thread structure are selected by a low GWP refrigerant heat exchange optimization design method; an evaporator interface tube is arranged on one side of the heat recovery heat exchange tube, the evaporator interface tube is a pipe without an elbow, and is connected to the evaporator inlet; a condenser interface tube is arranged on the other side of the heat recovery heat exchange tube, the condenser interface tube is a pipe with an elbow, and is connected to the condenser outlet; on the heat recovery heat exchange tube, a throttle valve interface is arranged between the condenser interface tube and the evaporator interface tube, and is used to connect to the throttle valve so that the refrigerant from the condenser flows to the throttle valve.

[0013] The advantages of the present invention are:

[0014] (1) The present invention first selects the type of refrigerant, determines the pre-set standard of the refrigeration system that meets the ecological environment requirements, and then selects the appropriate heat recovery heat exchange tube material and specifications, and optimizes the internal thread of the heat recovery heat exchange tube, so that the design accuracy of the heat recovery heat exchange tube is further improved and the heat exchange performance is improved.

[0015] (2) The present invention provides a basis for the internal thread parameters of the heat recovery heat exchange tube, ensuring that the heat recovery heat exchange tube can obtain the best heat exchange performance under different working conditions, providing a guarantee for improving the accuracy of the heat exchanger structure design calculation.

[0016] (3) The present invention first selects the type of refrigerant to determine the pre-set standards of the refrigeration system that meet the requirements of the ecological environment, then selects the appropriate heat recovery heat exchange tube material and specifications, and optimizes the internal threads of the heat recovery heat exchange tube. By considering the internal thread parameter value while calculating the heat recovery heat transfer coefficient on the tube side, the calculation of the heat transfer area of ​​the heat recovery heat exchange tube is made more accurate. By optimizing the internal thread structure of the heat recovery heat exchange tube, the heat exchange amount is increased and the heat exchange efficiency is improved, while also reducing the tube length, reducing the refrigerant pressure drop on the shell side, reducing energy waste, and saving manufacturing costs.

[0017] (4) The design method of the heat recovery heat exchange tube of the present invention breaks the design and technical bottleneck of traditional heat recovery heat exchange tubes, and provides users with more energy-saving, excellent, controllable, ecological, and heat exchange equipment with higher heat exchange performance to meet the higher demands of modern society for heat recovery heat exchange devices. BRIEF DESCRIPTION OF THE DRAWINGS

[0018] Figure 1 This is the working flow diagram of the refrigeration system;

[0019] Figure 2 This is the appearance diagram of the internally threaded heat recovery heat exchange tube;

[0020] Figure 3 It is a cross-sectional view of the internal threaded pipe structure;

[0021] Figure 4 It is a schematic diagram of the cross section of an internally threaded pipe;

[0022] Figure 5 Schematic diagram of the process of the present invention. DETAILED DESCRIPTION

[0023] The following will be combined with the drawings in the embodiments of the present invention to clearly and completely describe the technical solutions in the embodiments of the present invention. Obviously, the described embodiments are only part of the embodiments of the present invention, not all of the embodiments. Based on the embodiments of the present invention, all other embodiments obtained by ordinary technicians in this field without creative work are within the scope of protection of the present invention.

[0024] like Figure 1-2 As shown, the first aspect of the present invention proposes a regenerative heat exchange tube device, comprising: a regenerative heat exchange tube, a sleeve is arranged inside the regenerative heat exchange tube, the sleeve adopts an internal thread structure, a low GWP refrigerant heat exchange optimization design method is used to select the type of refrigerant and the GWP (global warming potential) is preliminarily selected for the internal thread structure parameters, and then the final value of the internal thread parameter is finally determined through iterative correction calculation; an evaporator interface tube is arranged on one side of the regenerative heat exchange tube, the evaporator interface tube is a small diameter pipe without an elbow, connected to the evaporator inlet, for the convenience of connection, the pipe length of the evaporator interface tube can be adjusted according to actual use conditions; a condenser interface tube is arranged on the other side of the regenerative heat exchange tube, the condenser interface tube is a small diameter pipe with an elbow, connected to the condenser outlet, for the convenience of connection, the elbow angle and the pipe length can be adjusted according to actual use conditions; on the regenerative heat exchange tube, a throttle valve interface is arranged between the condenser interface tube and the evaporator interface tube, for connecting with the throttle valve, so that the refrigerant from the condenser flows to the throttle valve.

[0025] The second aspect of the present invention provides a low GWP refrigerant heat exchange optimization design method, which is to design the heat exchange tube on the premise of achieving the rated standard carbon emissions. It includes:

[0026] S1: Select the type of refrigerant and determine the GWP of the refrigerant;

[0027] S2: Calculate the carbon footprint energy efficiency value G based on the basic data of the refrigerant and the heat exchange tube condenser side. The formula is:

[0028]

[0029] Among them, C is the carbon footprint value, EER is the energy efficiency value of the refrigeration system, Q L is the amount of refrigerant, S is the refrigerant leakage rate, GWP is the global warming potential of the refrigerant, which is determined according to the type of refrigerant, a is the operating years, and COP is the refrigeration coefficient, which is calculated by the calculation software.

[0030] The basic data on the condenser side include the condenser side inlet temperature t1, the subcooling degree △t1, and the condenser side outlet temperature t1'; wherein t1'=t1-△t1.

[0031] S3: Determine whether the carbon footprint value C and the refrigeration system energy efficiency value EER are compatible, that is, whether the carbon footprint energy efficiency value G meets the preset standard. If so, proceed to the next step. Otherwise, adjust the subcooling of the heat exchange tube. By adjusting the subcooling, the refrigeration system energy efficiency value EER is adjusted to make the carbon footprint energy efficiency value G meet the preset standard. That is, the carbon footprint value C and the refrigeration system energy efficiency value EER are compatible, ensuring the dynamic balance between the two, and then proceed to the next step.

[0032] Among them, the pre-set standards are refrigerants with GWP ranges of 0-150, 150-1500, and >1500, and the corresponding standard ranges of carbon footprint energy efficiency values ​​G are ≤10kg·a, ≤90kg·a, and ≤245kg·a, respectively.

[0033] S4: Based on the basic data of the condenser side and the evaporator side, estimate the heat transfer area and select appropriate evaporator side and condenser side pipe materials and specifications; including:

[0034] S41: Based on the basic data on the condenser side and the basic data on the evaporator side, determine the qualitative temperature t of the condenser side (shell side) of the heat exchange tube 1p and the qualitative temperature t of the evaporator side (tube side) 2p , obtain the physical parameters of the refrigerant fluid, as well as the heat transfer Q and the countercurrent logarithmic mean temperature difference △t m ,include:

[0035] Among them, the basic data on the condenser side include the inlet temperature t1 on the condenser side, the subcooling degree △t1, and the outlet temperature t1' on the condenser side; wherein t1'=t1-△t1; the basic data on the evaporator side include the inlet temperature t2 on the evaporator side, the outlet temperature t2' on the evaporator side, the superheat △t2, and the physical properties of the refrigerant fluid include the qualitative temperature on the condenser side and the evaporator side, the refrigerant specific heat on the condenser side and the evaporator side, the refrigerant dynamic viscosity on the condenser side and the evaporator side, the refrigerant thermal conductivity on the condenser side and the evaporator side, the refrigerant Prandtl number on the condenser side and the evaporator side, and the refrigerant density on the condenser side and the evaporator side.

[0036] S411: Calculate the qualitative temperature of the condenser side, the qualitative temperature of the condenser side t 1p The calculation formula is:

[0037]

[0038] S412: determining the physical property parameters of the refrigerant on the condenser side based on the qualitative temperature on the condenser side;

[0039] S413: Determine the heat transfer amount Q based on the specific heat in the physical property parameters. The heat transfer amount Q is calculated as follows:

[0040] Q=q1=q2=c p1 .q m .△t1

[0041] q1 is the heat exchange on the condenser side, q2 is the heat exchange on the evaporator side, c p1 is the specific heat of the refrigerant on the condenser side, q m is the refrigerant mass flow rate, obtained by calculation software;

[0042] S414: Obtain superheat Δt2 based on the heat exchange amount. c p2 is the specific heat of the refrigerant on the evaporator side, and the outlet temperature on the evaporator side is t2'=t2+△t2;

[0043] The qualitative temperature of the evaporator side needs to be assumed first. First, assume that the inlet temperature t2 of the evaporator side is the qualitative temperature of the evaporator side, and query the physical properties of the refrigerant on the evaporator side at this time; calculate the heat exchange q2 on the evaporator side according to the physical properties of the refrigerant, and Q = q1 = q2 = c p1 .q m .△t1=c p2 .q m .△t2 calculates the evaporator side outlet temperature t2'=t2+△t2; then calculates the average temperature of the evaporator side inlet temperature t2 and the evaporator side outlet temperature t2' as the actual qualitative temperature of the evaporator side The actual qualitative temperature t on the evaporator side 2pQuery and obtain the refrigerant physical properties parameters on the evaporator side.

[0044] S415: Obtaining the countercurrent logarithmic mean temperature difference △t m , △t max The larger value is the difference between the inlet temperature of the condenser and the outlet temperature of the evaporator, and the difference between the outlet temperature of the condenser and the inlet temperature of the evaporator. min It is the smaller value of the difference between the inlet temperature of the condenser side and the outlet temperature of the evaporator side and the difference between the outlet temperature of the condenser side and the inlet temperature of the evaporator side.

[0045] S42: Based on heat transfer Q and counterflow logarithmic mean temperature difference △t m Estimate the heat transfer area A, select the pipe material and specifications, and calculate the pressure drop △P on the evaporator side i And shell side pressure drop △P L .

[0046] The calculation formula for estimating the heat transfer area A is: Where K is the primary heat transfer coefficient.

[0047] Calculate the evaporator side pressure drop △P i The steps include:

[0048] S4211: Select the pipe material and specifications (outer diameter and wall thickness) on the evaporator side and calculate the required flow area on the evaporator side

[0049] Among them, d i is the inner diameter of the evaporator side tube, d i =d o -2*δ2,d o is the outer diameter of the evaporator side tube, δ2 is the wall thickness of the evaporator side tube;

[0050] S4212: Calculate the refrigerant flow rate in the heat exchange tube shell

[0051] S4213: Calculate the heat transfer area of ​​the structure A' = nπd o l; where n is the number of tubes, l is the estimated tube length, N p is the number of pipe passes.

[0052] S4214: Calculation of the Reynolds number on the evaporator side of the regenerative heat exchange tube Wherein, ρ2' is the refrigerant density on the evaporator side, and μ2' is the refrigerant dynamic viscosity on the evaporator side;

[0053] S4215: Calculate the tube side convection heat transfer coefficient h2, including:

[0054] judge Is it less than 60? If not, then Yes, then

[0055] Wherein, λ2' is the thermal conductivity of the refrigerant on the evaporator side, and Pr2' is the Prandtl number of the refrigerant on the evaporator side;

[0056] S4216: Calculate the evaporator side pressure drop △P i =△P L F t N p N t +△P n N t ; Among them, △P L is the pressure drop on the shell side; F t is the structural correction factor, N p is the number of tube passes, △P n is the pressure drop of the fluid flowing through the pipe inlet and outlet (Pa), N t is the number of shell passes in series.

[0057] Calculate the shell side pressure drop △P L The steps include:

[0058] S4221: Select the pipe material and specifications (shell outer diameter and shell wall thickness) on the shell side and calculate the shell flow cross-sectional area D i is the shell inner diameter, d o is the outer diameter of the tube on the tube side;

[0059] S4222: Calculation of the Reynolds number on the shell side Where, μ1 is the dynamic viscosity of the refrigerant on the condenser side, q m is the refrigerant mass flow rate;

[0060] S4223: Calculate the shell side heat transfer coefficient h1, including:

[0061] judge Is it less than 60? If not, then Yes, then

[0062] Where λ1 is the thermal conductivity of the refrigerant on the condenser side, d e is the equivalent diameter of the shell side, d e =D S -d O , D S is the outer diameter of the shell, Pr1 is the Prandtl number of the refrigerant on the condenser side, and l is the estimated tube length;

[0063] S4224: Calculate the shell side pressure drop Among them, λ o is the friction coefficient on the shell side, ρ1 is the refrigerant density on the condenser side, v1 is the refrigerant flow rate on the shell side,

[0064] S43: Determine whether the pressure drop on the evaporator side and the pressure drop on the shell side meet the standard conditions. If the pressure drop on the evaporator side does not meet the standard conditions, reselect the pipe material and specification on the evaporator side. If the pressure drop on the shell side does not meet the standard conditions, reselect the pipe material and specification on the shell side. If the pressure drops on both sides meet the standard conditions, proceed to the next step.

[0065] Among them, the standard condition is that the pressure drop on the condenser side and the evaporator side of the heat exchange tube must meet the standard pressure drop value established by the carbon footprint energy efficiency value, that is, the pressure drop on the shell side ≤JJ1, and the pressure drop on the evaporator side ≤JJ2.

[0066] In this embodiment, the pressure drop on the shell side needs to be ≤75 kPa, and the pressure drop on the evaporator side needs to be ≤25 kPa.

[0067] S44: Calculate the total heat transfer coefficient K' and the actual heat transfer area A of the heat exchange tube based on the heat transfer coefficient on the evaporator side and the shell side ” And the actual tube length l'; the calculation formula of the total heat transfer coefficient K' is:

[0068]

[0069] Actual heat transfer area A ” The calculation formula is:

[0070]

[0071] Among them, h2 is the heat transfer coefficient of the tube, d o is the outer diameter of the evaporator side tube, d i is the inner diameter of the evaporator side tube, λ t is the thermal conductivity of the pipe, h1 is the shell heat transfer coefficient; Q is the heat transfer amount, △t m is the countercurrent logarithmic mean temperature difference.

[0072] The calculation formula for the actual pipe length is: Where n is the number of tubes, N p is the number of pipe passes.

[0073] S45: Determine the actual heat transfer area A ” Does the heat transfer area A' satisfy k1·A"≤A'≤k2·A"? If yes, it proves that the above evaporator side structure and shell side structure are reasonably designed, and proceed to the subsequent steps; if no, jump to step S42; wherein, 0.9 <k1<k2<1.5;

[0074] In this embodiment, 1.05A"≤A'≤1.1A".

[0075] S5: Continue to optimize the internal thread of the heat recovery heat exchange tube to obtain the final optimized heat recovery heat exchange tube, including:

[0076] S51: Preliminary selection of the internal thread parameters of the heat recovery heat exchange tube, including wing height e, helix angle α, number of wing Ns, and pitch p;

[0077] S52: Based on the heat transfer coefficient h'2 of the evaporator side after the heat exchange tube is optimized, and in combination with the heat transfer coefficient of the shell side, the optimized total heat transfer coefficient K" and the optimized heat transfer area A"' are calculated;

[0078]

[0079] Among them, d o is the outer diameter of the evaporator side tube, λ t The thermal conductivity of the pipe, h1 is the heat transfer coefficient on the shell side, Q is the heat transfer, △t m is the countercurrent logarithmic mean temperature difference;

[0080] The heat transfer coefficient h'2 of the evaporator side after the heat recovery heat exchange tube is optimized includes:

[0081] judge Is it less than 60? If not, then yes, Where Re2 is the Reynolds number on the evaporator side, d i is the inner diameter of the tube on the evaporator side, λ2' is the thermal conductivity of the refrigerant on the evaporator side, and l is the estimated tube length;

[0082] S53: recalculate and obtain the optimized tube length l' based on the optimized heat transfer area, and determine whether the optimized tube length satisfies l'≤l' and l'=p×Ns. If yes, proceed to the next step; if no, readjust the internal thread parameters of the heat recovery heat exchange tube.

[0083] Among them, the calculation formula of the optimized tube length l" is:

[0084] When readjusting, the internal thread angle α is adjusted first, increasing by 1 degree each time until l” meets the requirement l”≤l', and then Ns and p are corrected, that is, l”= p× N s Until the design requirements are met;

[0085] S54: Calculate the optimized evaporator side friction coefficient f1 and the optimized shell side pressure drop △P' LAnd the optimized evaporator side pressure drop △P' i ;

[0086]

[0087] △P′ i =△P′ L F t N p N t +△P n N t

[0088] Where Re2 is the Reynolds number on the evaporator side, Ns' is the number of corrected rear fins, and λ o is the friction coefficient on the shell side, l” is the optimized tube length, d e is the equivalent diameter of the shell side, ρ1 is the refrigerant density on the condenser side, v1 is the refrigerant flow rate on the shell side, F t is the structural correction factor, N p is the number of tube passes, N t is the number of shell passes in series, △P n It is the pressure drop of the fluid flowing through the inlet and outlet of the pipeline.

[0089] S55: Determine whether the optimized shell-side pressure drop and the optimized evaporator-side pressure drop meet the standard conditions. If the optimized evaporator-side pressure drop or the optimized shell-side pressure drop does not meet the standard conditions, jump to step S51. If the pressure drops on both sides meet the standard conditions, the internal thread parameters of the heat recovery heat exchange tube at this time are used as the final optimization result to complete the final design of the heat recovery heat exchange tube.

[0090] The above-mentioned low GWP refrigerant heat exchange optimization design method is described below in conjunction with specific embodiments.

[0091] Example 1

[0092] Take the refrigerant type R404A (GWP is 3922) as an example.

[0093] 1. Calculation of carbon footprint energy efficiency value

[0094] Take the inlet temperature t1 of the condenser side as 40℃, the subcooling △t1 as 3℃, then the outlet temperature t1' of the shell side is 37℃, and the qualitative temperature of the condenser side is Calculate carbon footprint energy efficiency value The subcooling of the heat exchange tube needs to be adjusted. Taking the subcooling as 4°C, we get Then the subcooling of the heat exchange tube needs to be adjusted again, and the subcooling is set to 5℃, and we get The pre-set standard is met, so the supercooling is set to 5°C, and the other original data remain unchanged. Go to the next step of design.

[0095] 2. Raw data

[0096] Assuming the inlet temperature t1 of the condenser side (shell side) is 40°C and the subcooling △t1 is 5°C, the outlet temperature t1' of the shell side is 35°C.

[0097] In summary, we can get t1=40℃, △t1=5℃, t1'=35℃, and the qualitative temperature is The physical properties of the refrigerant fluid on the shell side are:

[0098] Specific heat of refrigerant on the condenser side c p1 =1683.9J / kg·K

[0099] Condenser side refrigerant dynamic viscosity μ1=0.000106p a ·s

[0100] Condenser side refrigerant thermal conductivity λ1=0.05767W / m·K

[0101] Condenser side refrigerant Prandtl number Pr1 = 3.11

[0102] Condenser side refrigerant density ρ1=979.1kg / m 3

[0103] Take the inlet temperature of the evaporator side (tube side) as 7°C. After the working conditions are determined, the refrigerant mass flow rate q in the heat exchanger can be calculated. m =0.295kg / s, assuming that t2=7℃ is taken as the qualitative temperature of the evaporator side, the assumed refrigerant physical properties of the evaporator side are:

[0104] Assumed specific heat of refrigerant on the evaporator side c p2 =1422.3J / kg·K

[0105] Assumed dynamic viscosity of refrigerant on the evaporator side μ2 = 0.0001645p a ·s

[0106] Assumed evaporator side refrigerant thermal conductivity λ2 = 0.07042 W / m·K

[0107] Assumed condenser side refrigerant Prandtl number Pr2 = 3.32

[0108] Assumed refrigerant density on the evaporator side ρ2 = 1122.9 kg / m 3

[0109] The heat transfer on the shell side can be calculated from the shell side data: q1 = c p1 .q m.△t1=2483.8W, from the heat transfer equation, it can be obtained that the heat transfer on the shell side is equal to the heat transfer on the evaporator side, that is, q1=q2=c p2 .q m .△t2=2483.8W, the outlet temperature of the evaporator side can be calculated as t2'=12.9℃, and △t2=5.9℃.

[0110] Pick is the qualitative temperature on the evaporator side, and the physical parameters of the refrigerant on the evaporator side are:

[0111] Specific heat of refrigerant on the evaporator side c p2 '=1437.5J / kg·K

[0112] Evaporator side refrigerant dynamic viscosity μ2'=0.000157p a ·s

[0113] Evaporator side refrigerant thermal conductivity λ2'=0.07W / m·K

[0114] Condenser side refrigerant Prandtl number Pr2'=3.3

[0115] Evaporator side refrigerant density ρ2'=1110.6kg / m 3 (Density on the evaporator side after correction)

[0116] From the above, we can get the countercurrent logarithmic mean temperature difference of this heat exchanger is △t m =27.55℃.

[0117] 3. Evaporator side calculation

[0118]

[0119] Select seamless copper tube Φ19×1, that is, the outer diameter of the pipe is do=0.019m, the wall thickness is δ2 is 0.001m, and the inner diameter is di=d o -2*δ2=0.017m

[0120] Required flow area on the tube side (evaporator side)

[0121] The refrigerant flow rate in the heat exchange tube is

[0122] Take the number of pipe passes N p =1, take the number of tubes per pass n = 1,

[0123] Estimated pipe length

[0124] Calculate the heat transfer area of ​​the structure

[0125] A'=nπd0l=1×3.14×0.019×0.76=0.0453㎡

[0126] Calculate the Reynolds number on the evaporator side of the regenerative heat exchange tube

[0127]

[0128] Calculate the tube side convection heat transfer coefficient h2

[0129] because When calculating the convective heat transfer coefficient on the evaporator side, the calculation formula needs to be multiplied by Make corrections, that is

[0130]

[0131] Calculation of pressure drop on evaporator side

[0132] The formula for calculating the pressure drop on the evaporator side is:

[0133] △P i =△P Li F t N p N t +△P n N t

[0134] In the formula, △P i is the pressure drop on the evaporator side (Pa); △P L F is the pressure drop caused by friction resistance when the fluid flows through a straight pipe (Pa); t For structural correction factors, take 1.5 for Φ19×1 pipe; N p is the number of tube passes; N t is the number of shell passes in series; △P n It is the pressure drop of the fluid flowing through the inlet and outlet of the pipeline (Pa).

[0135] Where △P L With △P n The calculation formula is

[0136]

[0137] In the formula, v2 is the refrigerant flow rate in the heat exchange tube shell in m / s; d i is the inner diameter of the tube; l is the estimated tube length; λ i is the friction coefficient, dimensionless, which can be obtained by looking up the graph; ρ2' is the density of the fluid in the tube kg / m 3 .

[0138] In summary, the pressure drop on the evaporator side is

[0139]

[0140] Meets the requirements.

[0141] 4. Shell side calculation

[0142] (1) Shell inner diameter

[0143] The shell on the shell side is Φ25×2, so its inner diameter D i =0.021m

[0144] Determine the aspect ratio When calculating the heat transfer coefficient, multiply by Correct the shell side flow cross-sectional area

[0145] The shell flow cross-sectional area on the shell side is

[0146] (3) Shell side Reynolds number

[0147]

[0148] (4) Calculation of pressure drop on the shell side

[0149] Since there is no baffle and the fluid flows along the tube, the shell side pressure drop is directly calculated according to the tube pressure drop formula when the fluid flows through the straight tube part, but the tube inner diameter in its characteristic size should be the equivalent diameter d e , d e =D S -d O =0.021-0.019=0.002m, and the shell side velocity

[0150] The pressure drop on the shell side is

[0151] Meet the standard requirements;

[0152] (5) Calculation of heat transfer coefficient on the shell side

[0153]

[0154]

[0155] 5. Calculation of heat transfer area required for heat exchange tubes

[0156] (1) Total heat transfer coefficient K' of the heat exchange tube

[0157]

[0158] (2) Calculation of actual heat transfer area

[0159] Therefore, it meets the requirements.

[0160] The actual tube length is

[0161] 6. Evaporator side optimization calculation

[0162] (1) Optimizing the internal thread;

[0163] (2) The inner tube is designed as a threaded type, and the Nusselt correlation is taken as

[0164]

[0165] Where e is the thread wing height m, d is the pipe inner diameter m, p is the pitch m, α is the helical angle, and Ns is the number of thread wings.

[0166] Known Right now (Refer to this formula to get the following formula)

[0167] The optimized heat transfer coefficient of the evaporator side can be obtained:

[0168]

[0169] Take e=0.0002m, α=5°, p=0.01m, Ns=25, is the correction factor, l' is the actual pipe length.

[0170] have to

[0171]

[0172] (3) Optimized overall heat transfer coefficient

[0173]

[0174] (4) Optimized heat transfer area

[0175]

[0176] Optimized tube length

[0177] Since p×Ns=0.01×25=0.25m≈0.243m, there is no need to correct Ns.

[0178] (5) Optimize the rear tube and shell resistance

[0179] Friction coefficient:

[0180]

[0181]

[0182] Meets standard requirements.

[0183] In accordance with the standard requirements, the design results of the heat recovery heat exchange tubes are output, and the calculation table is shown in Table 1.

[0184] Table 1 Design table of heat exchange tubes for refrigerant type R404A

[0185]

[0186]

[0187]

[0188]

[0189] Example 2

[0190] Select the type of refrigerant and determine the GWP value of the refrigerant, taking R448A (GWP is 1387) as an example.

[0191] 1. Calculation of carbon footprint energy efficiency value

[0192] If the inlet temperature of the condenser side is 40℃ and the subcooling degree is 3℃, the outlet temperature of the shell side is 37℃. That is, t1=40℃, △t1=3℃, t1'=37℃, and the qualitative temperature of the condenser side is Calculate carbon footprint energy efficiency value Therefore, there is no need to adjust the subcooling of the heat exchange tube, that is, take the subcooling as 3℃, and keep the other original data unchanged. Go to the next step of design.

[0193] 2. Raw data

[0194] If the inlet temperature of the condenser side (shell side) is 40°C and the subcooling degree is 3°C, the outlet temperature of the shell side is 37°C.

[0195] In summary, we can get t1s = 40 °C, △t1 = 3 °C, t1' = 37 °C, and the qualitative temperature is At this time, the working pressure on the shell side is 1.81MPa.

[0196] The physical properties of the refrigerant fluid on the shell side are:

[0197] Specific heat of refrigerant on the condenser side c p1 =1676.8J / kg·K

[0198] Condenser side refrigerant dynamic viscosity μ1=0.0001157p a ·s

[0199] Condenser side refrigerant thermal conductivity λ1=0.075W / m·K

[0200] Condenser side refrigerant Prandtl number Pr1 = 2.6

[0201] Condenser side refrigerant density ρ1=1032.9kg / m 3

[0202] Take the inlet temperature of the evaporator side (tube side) as 7°C. After the working conditions are determined, the refrigerant mass flow rate q in the heat exchanger can be calculated. m =0.326kg / s, assuming t2=7℃ as the qualitative temperature, the refrigerant physical properties on the tube side are:

[0203] Assumed specific heat of refrigerant on the evaporator side c p2 =1452J / kg·K

[0204] Assumed dynamic viscosity of refrigerant on the evaporator side μ2 = 0.000173p a ·s

[0205] Assumed evaporator side refrigerant thermal conductivity λ2 = 0.09 W / m·K

[0206] Assumed condenser side refrigerant Prandtl number Pr2 = 2.8

[0207] Assumed refrigerant density on the evaporator side ρ2 = 1171.3 kg / m 3

[0208] The heat transfer on the shell side can be calculated from the shell side data: q1 = c p1 .q m .△t1=1640W, from the heat transfer equation, it can be obtained that the heat transfer on the shell side is equal to the heat transfer on the tube side, that is, q1=q2=c p2 .q m .△t2=1640W, it can be calculated that the outlet temperature on the tube side is t2'=10.5℃, and △t2=3.5℃.

[0209] Pick is the qualitative temperature, and the working pressure on the tube side is 0.82MPa. The physical parameters of the refrigerant on the tube side are:

[0210] Specific heat of refrigerant on the evaporator side c p2 '=1460J / kg·K

[0211] Dynamic viscosity of refrigerant on the evaporator side μ2'=0.00017p a ·s

[0212] Evaporator side refrigerant thermal conductivity λ2'=0.089W / m·K

[0213] Condenser side refrigerant Prandtl number Pr2'=2.78

[0214] Evaporator side refrigerant density ρ2'=1164.5kg / m 3

[0215] From the above, we can get the countercurrent logarithmic mean temperature difference of this heat exchanger is △t m =29.75℃.

[0216] 3. Calculation on the pipe side

[0217] (1) According to the data, the heat transfer coefficient K is estimated to be 2000W / ㎡·K, and the heat transfer area of ​​the heat exchange tube is estimated to be

[0218]

[0219] Select seamless copper tube Φ19×1, that is, the outer diameter of the pipe is do=0.019m, the wall thickness is δ2 is 0.001m, and the inner diameter is di=d o -2*δ2=0.017m

[0220] Required flow area on the tube side

[0221] The refrigerant flow rate in the heat exchange tube is

[0222] Take the number of pipe passes N p =1, the number of tubes per pass n = 1,

[0223] Estimated pipe length

[0224] Calculate the heat transfer area of ​​the tube

[0225] A'=nπd0l=1×3.14×0.019×0.47=0.028㎡

[0226] Calculate the Reynolds number on the evaporator side of the regenerative heat exchange tube

[0227]

[0228] Calculate the tube side convection heat transfer coefficient h2

[0229] because When calculating the convective heat transfer coefficient on the tube side, the calculation formula needs to be multiplied by Make corrections, that is

[0230]

[0231] Calculation of pressure drop on the tube side

[0232] The formula for calculating the pressure drop on the tube side is:

[0233] △P i =△P Li F t N p N t +△P n N t

[0234] In the formula, △P i is the pressure drop on the tube side (Pa); △P L F is the pressure drop caused by friction resistance when the fluid flows through a straight pipe (Pa); t For structural correction factors, take 1.5 for Φ19×1 pipe; N p is the number of tube passes; N t is the number of shell passes in series; △P n It is the pressure drop (Pa) of the fluid flowing through the inlet and outlet of the pipe box.

[0235] Where △P L With △P n The calculation formula is

[0236]

[0237] In the formula, v2 is the refrigerant flow rate in the heat exchange tube shell in m / s; d i is the inner diameter of the tube; l is the estimated tube length; λ i is the friction coefficient, dimensionless, which can be obtained by looking up the graph; ρ2' is the density of the fluid in the tube kg / m 3 .

[0238] In summary, the pressure drop on the tube side is

[0239]

[0240] Meets the requirements.

[0241] 4. Shell side calculation

[0242] (1) Shell inner diameter

[0243] The shell on the shell side is Φ25×2, so its inner diameter D i =0.021m

[0244] Determine the aspect ratio When calculating the heat transfer coefficient, multiply by Correct the shell side flow cross-sectional area

[0245] The shell flow cross-sectional area on the shell side is

[0246] (3) Shell side Reynolds number

[0247]

[0248] (4) Calculation of pressure drop on the shell side

[0249] Since there is no baffle and the fluid flows along the tube, the shell side pressure drop is directly calculated according to the tube pressure drop formula when the fluid flows through the straight tube part, but the tube inner diameter in its characteristic size should be the equivalent diameter d e , d e =D S -d O =0.021-0.019=0.002m, and the shell side velocity

[0250] The pressure drop on the shell side is

[0251] Meet the standard requirements;

[0252] (5) Calculation of heat transfer coefficient on the shell side

[0253]

[0254] 5. Calculation of heat transfer area required for heat exchange tubes

[0255] (1) Total heat transfer coefficient of heat exchange tube

[0256]

[0257] (2) Actual heat transfer area

[0258] Therefore, it meets the requirements.

[0259] The actual tube length is

[0260] 6. Optimization calculation of pipe side

[0261] (1) Optimizing the internal thread;

[0262] (2) The inner tube is designed as a threaded type, and the Nusselt correlation is taken as

[0263]

[0264] Where e is the spiral wing height m, d is the inner diameter of the pipe m, p is the pitch m, α is the helical angle, and Ns is the number of spiral wings.

[0265] Known Right now

[0266] Available

[0267]

[0268] Take e=0.0002m, α=5°, p=0.01m, Ns=25,

[0269] have to

[0270]

[0271] (3) Optimized overall heat transfer coefficient

[0272]

[0273] (4) Optimized heat transfer area

[0274] Meets the requirements.

[0275] Optimized tube length

[0276] Correction: Since p×Ns=0.01×25=0.25m>0.17m, it is necessary to correct Ns, taking Ns′=17,

[0277] Then

[0278]

[0279] Meets the requirements.

[0280] Meets the requirements, so no correction is required

[0281] (5) Optimize the rear tube and shell resistance

[0282] Friction coefficient:

[0283]

[0284]

[0285] Meets standard requirements.

[0286] In accordance with the standard requirements, the design results of the heat recovery heat exchange tubes are output, and the calculation table is shown in Table 2 below.

[0287] Table 2 Design table of heat exchange tubes for refrigerant type R448A

[0288]

[0289]

[0290]

[0291]

[0292]

[0293] Of course, it is obvious to those skilled in the art that the present invention is not limited to the details of the exemplary embodiments described above, but also includes the same or similar structures that can be implemented in other specific forms without departing from the spirit or essential features of the present invention. Therefore, no matter from which point of view, the embodiments should be regarded as exemplary and non-limiting, and the scope of the present invention is defined by the appended claims rather than the above description, and it is intended that all changes falling within the meaning and scope of the equivalent elements of the claims are included in the present invention. Any reference numerals in the claims should not be regarded as limiting the claims involved.

[0294] In addition, it should be understood that although the present specification is described according to implementation modes, not every implementation mode contains only one independent technical solution. This description of the specification is only for the sake of clarity. Those skilled in the art should regard the specification as a whole. The technical solutions in each embodiment may also be appropriately combined to form other implementation modes that can be understood by those skilled in the art.

[0295] The techniques, shapes, and structural parts not described in detail in the present invention are all well-known techniques.

Claims

1. A low GWP refrigerant heat exchange optimization design method, characterized in that: include: S1: Select the type of refrigerant and determine the GWP of the refrigerant; GWP is the global warming potential; S2: Calculate the carbon footprint energy efficiency value G based on the GWP of the refrigerant; S3: Determine whether the carbon footprint value and the energy efficiency value of the refrigeration system are compatible, that is, whether the carbon footprint energy efficiency value meets the preset standard. If yes, proceed to the next step; otherwise, adjust the subcooling degree of the heat recovery heat exchange tube so that the carbon footprint energy efficiency value meets the preset standard, and then proceed to the next step; S4: Based on the basic data of the condenser side and the evaporator side, estimate the heat transfer area and select appropriate pipe materials and specifications for the evaporator side and the condenser side; S5: Continue to optimize the internal thread design of the heat recovery heat exchange tube to determine whether the optimized tube length and the optimized pressure drop on both sides meet the standard conditions. If yes, obtain the final design result of the heat recovery heat exchange tube. If not, re-evaluate the internal thread parameters of the heat recovery heat exchange tube and continue to optimize the design.

2. A low GWP refrigerant heat exchange optimization design method according to claim 1, characterized in that: In step S5, the internal thread of the heat recovery heat exchange tube is further optimized and designed to determine whether the optimized tube length and the optimized pressure drop on both sides meet the standard conditions, including: S51: Preliminary selection of the internal thread parameters of the heat recovery heat exchange tube, including wing height e, helix angle α, number of wing Ns, and pitch p; S52: Calculate the optimized total heat transfer coefficient K" and the optimized heat transfer area A"' based on the heat transfer coefficient h'2 of the evaporator side after the heat exchange tube is optimized and combined with the heat transfer coefficient of the condenser side; S53: recalculate and obtain the optimized tube length l' based on the optimized heat transfer area, and determine whether the optimized tube length satisfies l'≤l' and l'=p×Ns. If yes, proceed to the next step; if no, readjust the internal thread parameters of the heat recovery heat exchange tube; wherein l' is the actual tube length; S54: Calculate the optimized evaporator side friction coefficient f1 and the optimized condenser side pressure drop △P' L And the optimized evaporator side pressure drop △P' i ; S55: Determine whether the optimized condenser side pressure drop and the optimized evaporator side pressure drop meet the standard conditions. If the optimized evaporator side pressure drop or the optimized condenser side pressure drop does not meet the standard conditions, jump to step S51. If the pressure drops on both sides meet the standard conditions, the internal thread parameters of the heat recovery heat exchange tube at this time are used as the final optimization results to complete the final design of the heat recovery heat exchange tube.

3. A low GWP refrigerant heat exchange optimization design method according to claim 1, characterized in that: The carbon footprint energy efficiency value G is calculated based on the GWP of the refrigerant. The formula is: Among them, C is the carbon footprint value, EER is the energy efficiency value of the refrigeration system, Q L is the amount of refrigerant used, S is the refrigerant leakage rate, GWP is the global warming potential of the refrigerant, which is determined according to the type of refrigerant, a is the operating years, and COP is the refrigeration coefficient.

4. A low GWP refrigerant heat exchange optimization design method according to claim 1, characterized in that: In step S3, the pre-set standard is as follows: If the GWP value range is 0-150, the standard range of the carbon footprint energy efficiency value G is ≤10kg·a; If the GWP value is 150-1500, the standard range of the carbon footprint energy efficiency value G is ≤90kg·a; If the GWP value is >1500, the standard range of the carbon footprint energy efficiency value G is ≤245kg·a.

5. A low GWP refrigerant heat exchange optimization design method as claimed in claim 2, characterized in that: In step S4, the heat transfer area is estimated based on the basic data of the condenser side and the basic data of the evaporator side, and appropriate pipe materials and specifications of the evaporator side and the condenser side are selected, including: S41: Based on the basic data on the condenser side and the basic data on the evaporator side, determine the qualitative temperature t of the heat exchange tube condenser side. 1p and the evaporator side qualitative temperature t 2p , obtain the physical properties of the refrigerant fluid, as well as the heat transfer Q and the countercurrent logarithmic mean temperature difference △t m ; S42: Estimate the heat transfer area based on the heat transfer and countercurrent logarithmic mean temperature difference, select the pipe material and specifications, and calculate the evaporator side pressure drop △P i and condenser side pressure drop △P L ; S43: Determine whether the pressure drop on the evaporator side and the pressure drop on the condenser side meet the standard conditions. If the pressure drop on the evaporator side does not meet the standard conditions, reselect the pipe material and specifications on the evaporator side. If the pressure drop on the condenser side does not meet the standard conditions, reselect the pipe material and specifications on the condenser side. If the pressure drops on both sides meet the standard conditions, proceed to the next step. S44: Calculate the total heat transfer coefficient K', the actual heat transfer area A'' and the actual tube length l' of the heat recovery heat exchange tube based on the heat transfer coefficient on the evaporator side and the heat transfer coefficient on the condenser side; S45: Determine whether the actual heat transfer area A" and the structural heat transfer area A' meet the basic conditions; if yes, it proves that the above evaporator side structure and condenser side structure are reasonably designed, and proceed to the subsequent step S5; if no, jump to step S42.

6. A low GWP refrigerant heat exchange optimization design method as claimed in claim 5, characterized in that: In step S41, the basic data on the condenser side include the inlet temperature t1 on the condenser side, the subcooling degree Δt1, and the outlet temperature t1' on the condenser side; wherein t1'=t1-Δt1; the basic data on the evaporator side include the inlet temperature t2 on the evaporator side, the outlet temperature t2' on the evaporator side, and the superheat degree Δt2; the physical properties of the refrigerant fluid include the refrigerant specific heat on the condenser side and the evaporator side, the refrigerant dynamic viscosity on the condenser side and the evaporator side, the refrigerant thermal conductivity on the condenser side and the evaporator side, the refrigerant Prandtl number on the condenser side and the evaporator side, and the refrigerant density on the condenser side and the evaporator side; In step S43, the standard condition is that the pressure drop on the condenser side and the evaporator side of the regenerative heat exchange tube must meet the standard pressure drop value established by the carbon footprint energy efficiency value, that is, the pressure drop on the condenser side ≤ JJ1, the pressure drop on the evaporator side ≤ JJ2, JJ1 and JJ2 are the pressure drop thresholds on the condenser side and the evaporator side respectively; In step S45, the basic condition is: k1·A”≤A'≤k2·A”, where 0.9 <k1<k2<1.5。 7. A low GWP refrigerant heat exchange optimization design method according to claim 6, characterized in that: In step S41, the condenser side temperature t 1p and the evaporator side qualitative temperature t 2p are the average values ​​of the inlet temperature on both sides and the outlet temperature on both sides respectively; The calculation formula for heat transfer Q is: Q=q1=q2=c p1 .q m .△t1 Among them, q1 is the heat exchange on the condenser side, q2 is the heat exchange on the evaporator side, c p1 is the specific heat of the refrigerant on the condenser side, q m is the refrigerant mass flow rate; Countercurrent logarithmic mean temperature difference △t m The calculation formula is: Among them, △t max The larger value is the difference between the inlet temperature of the condenser and the outlet temperature of the evaporator, and the difference between the outlet temperature of the condenser and the inlet temperature of the evaporator. min The smaller value is the difference between the inlet temperature of the condenser side and the outlet temperature of the evaporator side and the difference between the outlet temperature of the condenser side and the inlet temperature of the evaporator side. In step S42, the calculation formula for estimating the heat transfer area A is: Where K is the primary heat transfer coefficient; Calculate the evaporator side pressure drop △P i The steps include: S4211: Select the pipe material and specifications on the evaporator side and calculate the required flow area on the evaporator side Among them, d i is the inner diameter of the evaporator side tube, d i =d o -2*δ2,d o is the outer diameter of the evaporator side tube, δ2 is the wall thickness of the evaporator side tube; S4212: Calculate the refrigerant flow rate in the heat exchange tube shell S4213: Calculate the heat transfer area of ​​the structure A' = nπd o l; where n is the number of tubes, l is the estimated tube length, N p is the number of pipe passes; S4214: Calculation of the Reynolds number on the evaporator side of the regenerative heat exchange tube Wherein, ρ2' is the refrigerant density on the evaporator side, and μ2' is the refrigerant dynamic viscosity on the evaporator side; S4215: Calculate the evaporator side convection heat transfer coefficient h2, including: judge Is it less than the set threshold m1? If not, then Yes, then Wherein, λ2' is the thermal conductivity of the refrigerant on the evaporator side, and Pr2' is the Prandtl number of the refrigerant on the evaporator side; S4216: Calculate the evaporator side pressure drop △P i =△P L F t N p N t +△P n N t ; Among them, △P L is the pressure drop on the condenser side; F t is the structural correction factor, N p is the number of tube passes, △P n is the pressure drop of the fluid flowing through the pipe inlet and outlet (Pa), N t is the number of shell passes in series; Calculate the pressure drop △P on the condenser side L The steps include: S4221: Select the pipe material and specifications on the condenser side and calculate the shell flow cross-sectional area D i is the shell inner diameter, d o is the outer diameter of the evaporator side tube; S4222: Calculate the Reynolds number on the condenser side Where, μ1 is the dynamic viscosity of the refrigerant on the condenser side, q m is the refrigerant mass flow rate; S4223: Calculate the heat transfer coefficient h1 on the condenser side, including: judge Is it less than the set threshold m1? If not, then Yes, then h1= Where λ1 is the thermal conductivity of the refrigerant on the condenser side, d e is the equivalent diameter of the condenser side, d e =D S -d O , D S is the outer diameter of the shell, Pr1 is the Prandtl number of the refrigerant on the condenser side, and l is the estimated tube length; S4224: Calculate the pressure drop on the condenser side Among them, λ o is the friction coefficient on the condenser side, ρ1 is the refrigerant density on the condenser side, v1 is the refrigerant flow rate on the condenser side, 8. A low GWP refrigerant heat exchange optimization design method as claimed in claim 5, characterized in that: In step S44, the calculation formula of the total heat transfer coefficient K' is: The calculation formula of the actual heat transfer area A" is: The calculation formula of the actual pipe length l' is: Where h2 is the heat transfer coefficient on the evaporator side, d o is the outer diameter of the evaporator side tube, d i is the inner diameter of the evaporator side tube, λ t is the thermal conductivity of the pipe, h1 is the heat transfer coefficient on the condenser side; Q is the heat transfer amount, △t m is the countercurrent logarithmic mean temperature difference, n is the number of tubes, N p is the number of pipe passes.

9. A low GWP refrigerant heat exchange optimization design method according to claim 7, characterized in that: In the step S52, the heat transfer coefficient h'2 of the evaporator side after the heat recovery heat exchange tube is optimized is obtained, including: judge Is it less than 60? If not, then yes, Where Re2 is the Reynolds number on the evaporator side, d i is the inner diameter of the tube on the evaporator side, λ2' is the thermal conductivity of the refrigerant on the evaporator side, and l is the estimated tube length; The calculation formula of the optimized heat transfer area A'' is: Where K” is the optimized total heat transfer coefficient, d o is the outer diameter of the evaporator side tube, λ t is the thermal conductivity of the pipe, h1 is the heat transfer coefficient on the condenser side, Q is the heat transfer amount, △t m is the countercurrent logarithmic mean temperature difference, h'2 is the heat transfer coefficient on the evaporator side after the heat recovery heat exchange tube is optimized; In step S53, the calculation formula of the optimized tube length is: In step S54, the friction coefficient f1 on the evaporator side and the pressure drop ΔP' on the condenser side are optimized. L And the optimized evaporator side pressure drop △P' i The calculation formula is: Where Re2 is the Reynolds number on the evaporator side, Ns' is the number of corrected rear fins, and λ o is the friction coefficient on the condenser side, l” is the optimized tube length, d e is the equivalent diameter of the shell side, ρ1 is the refrigerant density on the condenser side, v1 is the refrigerant flow rate on the shell side, F t is the structural correction factor, N p is the number of tube passes, N t is the number of shell passes in series, △P n It is the pressure drop of the fluid flowing through the inlet and outlet of the pipeline.

10. A heat regeneration heat exchange tube device, characterized in that: include: A heat recovery heat exchange tube, wherein a sleeve is arranged inside the heat recovery heat exchange tube, and the sleeve adopts an internal thread structure. According to a low GWP refrigerant heat exchange optimization design method shown in any one of claims 1 to 9, the type of refrigerant and the internal thread structure are selected; an evaporator interface pipe is arranged on one side of the heat recovery heat exchange tube, and the evaporator interface pipe is a pipe without an elbow, which is connected to the evaporator inlet; a condenser interface pipe is arranged on the other side of the heat recovery heat exchange tube, and the condenser interface pipe is a pipe with an elbow, which is connected to the condenser outlet; on the heat recovery heat exchange tube, a throttle valve interface is arranged between the condenser interface pipe and the evaporator interface pipe, which is used to connect to the throttle valve so that the refrigerant from the condenser flows to the throttle valve.

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