JP2012122692A - Heat transfer tube with grooved inner surface - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a heat transfer tube with a grooved inner surface, which has high evaporation performance even when heat pump equipment using a COrefrigerant is used under a low heat flux (q) condition, and polyalkylene glycol-based refrigerating machine oil is mixed in the refrigerant, and which reduces failure of a compressor by improving oil return.SOLUTION: The heat transfer tube 20 with a grooved inner surface is used for an evaporator provided in a heat pump water heater that uses carbon dioxide as a refrigerant containing 0.1 mol% or more of polyalkylene glycol-based refrigerating machine oil, and that is used at a heat flux (q) of 10 kW/m. The heat transfer tube 20 with the grooved inner surface includes a main tube 21 which includes a plurality of spiral grooves 22 arranged in parallel to each other on the inner peripheral surface and a plurality of fins 23 arranged between the plurality of grooves 22. The inner diameter (D) of the main tube 21 is 2.4 to 6.8 mm, the height (HF) of the fin 23 is 0.1 to 0.25 mm, and 0.04≤W/ID≤0.1 is satisfied when Wis a bottom width of the plurality of grooves 22.

Description

  The present invention relates to an internally grooved heat transfer tube. In particular, the present invention relates to an internally grooved heat transfer tube using carbon dioxide as a refrigerant.

Conventionally, chlorofluorocarbon refrigerants have been used in heat exchangers provided in air conditioners, car air conditioners, refrigerators, refrigerators, water heaters, vending machines, and the like. However, chlorofluorocarbon-based refrigerants are feared to have an impact on global warming, and are carbon dioxide, a natural refrigerant that is safe and inexpensive without toxicity and flammability, has low environmental impact, and has excellent thermal properties. (CO ₂ ) is drawing attention. Further, in Patent Document 1, since carbon dioxide has a small surface tension, bubbles are more likely to be generated than chlorofluorocarbon refrigerants, and nucleate boiling is promoted. Therefore, when carbon dioxide is used as a refrigerant, a fluorocarbon refrigerant is used. It is described that the heat transfer performance is improved as compared with the above.

On the other hand, as a method for improving the evaporation performance of the heat transfer tube, there is a method of expanding the heat transfer area by forming a groove on the inner surface to solve the problem that the theoretical performance in the simple cycle of cooling and heating of the CO ₂ refrigerant is low. However, the CO ₂ refrigerant that circulates in the heat transfer pipe contains refrigeration oil that is a lubricant for the compressor, and the inner grooved pipe used in the chlorofluorocarbon refrigerant is simply used for the CO ₂ refrigerant. Even if it is used, sufficient heat transfer performance cannot be obtained.

  In view of such problems, Patent Document 1 discloses an inner grooved heat transfer tube for an evaporator that does not increase pressure loss and has excellent evaporation heat transfer performance even when a carbon dioxide refrigerant containing refrigeration oil is used. providing.

Non-Patent Document 1 describes the evaporation performance of an internally grooved tube, and although evaluation as a single tube is carried out at 25 kW / m ² or more, a heat pump hot water supply that actually uses CO ₂ refrigerant. The machine describes that the heat flux (q) is as small as about 5 kW / m ^{2 in the} rated specification. Since the influence of nucleate boiling is overestimated depending on the test conditions, it is considered that there is a possibility that an optimum fin shape may not be obtained under the low heat flux (q) condition such as a CO ₂ refrigerant heat pump water heater.

Japanese Patent No. 4386813

Li Sobu, Saeki Main Tax and three others, "Experimental Study on Evaporative Heat Transfer Enhancement of Natural Refrigerant CO2 in Pipes" Copper and Copper Alloy Technology Workshop (November 12 and 13, 2004), 44th Lecture Meeting Overview (p79, 80)

As described above, Patent Document 1 shows that nucleate boiling is promoted as a characteristic of the CO ₂ refrigerant. In general, although the effect of nucleate boiling increases as the heat flux (q) increases, Patent Document 1 does not describe the influence of the heat flux (q), and considers the heat flux (q) condition on the basis of numerical limitation. It has not been.

In the CO ₂ refrigerant heat pump water heater described in Non-Patent Document 1, polyalkylene glycol (PAG) that is highly stable at high temperature is used as the refrigeration oil. However, since PAG has low compatibility with CO ₂ and oil return is poor, more refrigeration oil stays in the heat transfer tube, and the presence of the oil film greatly affects the suppression of nucleate boiling. If the oil return is poor, the lubrication of the compressor will be poor and cause failure, so it is conceivable to install an oil separator on the discharge side of the compressor as a preventive measure. However, there is a problem of cost increase and space, which is not practical in a household CO ₂ refrigerant heat pump water heater, and more PAG is actually enclosed so that oil does not disappear from the compressor.

Therefore, an object of the present invention is to have an excellent evaporation performance even when a heat pump device using a CO ₂ refrigerant is used under a low heat flux (q) condition and a polyalkylene glycol refrigerating machine oil is mixed in the refrigerant. An object of the present invention is to provide an internally grooved heat transfer tube that can reduce the risk of a compressor failure by improving oil return.

In order to achieve the above object, the present invention uses carbon dioxide having a content of polyalkylene glycol refrigerating machine oil of 0.1% by mass or more as a refrigerant, and has a heat flux (q) of less than 10 kW / m ² . A heat transfer tube with an inner surface groove used in an evaporator provided in the heat pump water heater used, and a plurality of spiral grooves provided in parallel to each other on an inner peripheral surface, and a plurality of fins positioned between the plurality of grooves The inner diameter (ID) of the main pipe is 2.4 mm or more and 6.8 mm or less, and the height (HF) of the fin is 0.1 mm or more and 0.25 mm or less. An inner-grooved heat transfer tube that satisfies 0.04 ≦ W ₃ /ID≦0.1 is provided where the width of the bottom of the groove is W ₃ .

  Further, in the inner surface grooved heat transfer tube, an angle formed between the groove and the fin may be 15 ° or more and 25 ° or less.

  Further, in the inner surface grooved heat transfer tube, the plurality of fins may be 30 or more and 50 or less.

According to the internally grooved heat transfer tube according to the present invention, the heat pump device using the CO ₂ refrigerant is used under a low heat flux (q) condition, and is excellent even when the polyalkylene glycol type refrigeration oil is mixed in the refrigerant. It is possible to provide an internally grooved heat transfer tube that has evaporation performance and can reduce the risk of compressor failure by improving oil return.

It is a schematic diagram of the structure of the heat pump water heater comprised using the inner surface grooved heat exchanger tube which concerns on embodiment of this invention. (A) And (b) is sectional drawing perpendicular | vertical to the tube-axis direction of the heat exchanger tube with an inner surface groove | channel which concerns on this Embodiment. It is a schematic diagram at the time of cutting out a part of internally grooved heat exchanger tube along a tube axis. It is a figure which shows the relationship between the internal diameter and wall thickness which become a pressure | voltage resistant equivalent to the inner surface grooved pipe | tube which consists of phosphorus deoxidized copper of outer diameter 7mm and bottom wall thickness 0.45mm (inner diameter 6.1mm). It is sectional drawing of schematic structure of the internally grooved tube processing apparatus for demonstrating the processing method of an internally grooved heat exchanger tube. Is an external view of the evaporator shunts are actually used CO ₂ refrigerant heat pump water heater. (A) is sectional drawing of the shunt part of a shunt with the outer diameter of a heat exchanger tube being 7 mm, (b) is sectional drawing assumed of the shunt with the outer diameter of a heat exchanger tube being 3 mm. It is a schematic diagram at the time of inserting the copper pipe before and after fractionation into a fractionator. It is a schematic diagram of a two-row 30-stage two-pass heat exchanger. It is a figure which shows the result of having calculated the pressure loss ratio with respect to 7 mm of outer diameters (inner diameter of 6.1 mm) x 2 path | passes when the inside diameter is made into a parameter and the number of passes is 10. Calculation result of fin height and inner surface area smooth tube ratio for heat transfer tube with inner diameter of 6.8mm and inner diameter of 2.4mm when the ratio of groove bottom width to inner diameter (W ₃ / ID) is constant at 0.04 FIG. It is a figure which shows the result of having measured the rotation speed ratio of a compressor, and the mixing amount to the refrigerant | coolant of refrigeration oil. The inner diameter and the fin height is 6.8 mm, if the 0.25mm and 2.4 mm, for the case of 0.1 mm, with the same _fin-shaped, a diagram showing the results of calculating the W 3 / ID and the inner surface smooth tube ratio is there. It is the schematic of a measuring apparatus. It is a schematic diagram of a heat exchanger tube measurement part (test section). It is a figure which shows a test result as a relationship between the number of fins and a heat transfer rate smooth tube ratio. W 3 _/ ID is a diagram showing the relationship between the heat transfer coefficient smooth tube ratio / single heavy smooth tube ratio.

[Embodiment]
FIG. 1: shows the outline | summary of a structure of the heat pump water heater comprised using the heat transfer tube with an inner surface groove | channel which concerns on embodiment of this invention.

  The heat pump water heater is connected to the compressor 11, the gas cooler 12 connected to the compressor 11 via piping, the expansion valve 13 connected to the gas cooler 12 via piping, and the expansion valve 13 via piping. The evaporator 14 is provided. The evaporator 14 and the compressor 11 are also connected to each other through piping. A compressor 11, a gas cooler 12, an expansion valve 13, an evaporator 14, and piping constitute a refrigeration cycle. In addition, carbon dioxide refrigerant, which is a refrigerant, is enclosed in the refrigeration cycle. The carbon dioxide refrigerant is in a supercritical state on the heat radiation side (that is, from the discharge part of the compressor 11 to the inlet part of the expansion valve 13 through the gas cooler 12). 1 is used for the gas cooler 12 and / or the evaporator 14.

  The outline of the operation of the heat pump water heater is as follows.

First, the carbon dioxide refrigerant (CO ₂ refrigerant) compressed by the compressor 11 enters a high-temperature and high-pressure state and is introduced into the gas cooler 12. At this time, the refrigerating machine oil which is the lubricating oil of the compressor 11 is also introduced into the gas cooler 12 simultaneously with the CO ₂ refrigerant. In the gas cooler 12, the CO ₂ refrigerant in the supercritical state does not enter a gas-liquid two-phase state, and the heat of the CO ₂ refrigerant is radiated to water or the like. Next, the CO ₂ refrigerant is decompressed by the expansion valve 13 to be in a low-pressure gas-liquid two-phase state, and is introduced into the evaporator 14 in this state. In the evaporator 14, the CO ₂ refrigerant absorbs heat from the air to become a gas state and is sucked into the compressor 11 again.

By repeating such a cycle, the gas cooler 12 is heated by heat dissipation and the evaporator 14 is cooled by heat absorption. Incidentally, polyalkylene glycol refrigerating machine oil used in the CO ₂ refrigerant, but also in combination with prone CO ₂ refrigerant to a high temperature has a feature such as excellent thermal stability, compatibility with the CO ₂ refrigerant is poor When the CO ₂ refrigerant evaporates from the refrigerant, it is easy to separate into three phases like a refrigerant gas phase, a refrigerant liquid phase, and refrigeration oil. Since it is easy to separate three phases, an oil film of polyalkylene glycol refrigerating machine oil is formed, and this oil film becomes a heat resistance. Therefore, if the refrigerating machine oil is mixed even a little, the evaporation performance is lowered. Further, the evaporator 14 is configured, for example, by arranging two rows and 30 stages of internally grooved heat transfer tubes having an inner diameter of 6.1 mm and a length of 800 mm. In this case, since the capacity of the evaporator of the CO ₂ refrigerant heat pump water heater having a rated capacity of 6 kW is about 4.6 kW, the heat flux (q) is about 5 kW / m ² . In the case of a model with a rated capacity of 4.5 kW, the capacity of the evaporator is smaller, so that the heat flux (q) is also smaller.

  FIGS. 2A and 2B show an outline of a cross section perpendicular to the tube axis direction of the internally grooved heat transfer tube according to the present embodiment. Specifically, (a) in FIG. 2A shows an outline of the cross section of the internally grooved heat transfer tube, and (b) in FIG. 2A is a partial cross section in which the region surrounded by the one-dot chain line in (a) in FIG. 2A is enlarged. Indicates. FIG. 2B shows an outline when a part of the internally grooved heat transfer tube is cut along the tube axis.

As shown in FIG. 2A (a) and FIG. 2B, the inner surface grooved heat transfer tube 20 according to the present embodiment includes a plurality of spiral grooves 22 provided on the inner peripheral surface in parallel at regular intervals, A main pipe 21 having a plurality of fins 23 having a mountain-like cross section located between the plurality of grooves 22 is provided. A groove 22 is formed between one fin 23 and another fin 23 adjacent to the one fin 23. In the present embodiment, the inner diameter (ID) of the main pipe 21 is 2.4 mm or more and 6.8 mm or less, and the fin height (HF), that is, the depth of the groove 22 is 0.1 mm or more and 0.25 mm. The following is satisfied, and 0.04 ≦ W ₃ /ID≦0.1 is satisfied when the width of the bottom of the plurality of grooves is W ₃ . The main pipe 21 is formed using a metal material such as copper or a copper alloy, for example.

Further, when the outer diameter of the inner grooved tube 20 is “OD” and the thickness of the main tube 21 is “TW”, the inner diameter “ID” can be expressed by “ID = OD−2 × TW”. The root width of the fin 23, that is, the root width of the fin 23 at the contact portion between the fin 23 and the inner peripheral surface of the main pipe 21 is defined as “W ₁ ”. Still referring to FIG. In this embodiment, an angle formed by the groove 22 and the fin 23 with respect to the longitudinal direction of the straight inner-surface grooved heat transfer tube 20 (hereinafter referred to as “twist angle”) is “β”.

The internally grooved heat transfer tube 20 according to the present embodiment can be used, for example, in an evaporator provided in a heat pump water heater. In addition, this heat pump water heater uses carbon dioxide having a polyalkylene glycol refrigerating machine oil content of 0.1% by mass or more as a refrigerant and a heat flux (q) of less than 10 kW / m ^2. it can.

(Regarding the upper limit of the ID of the heat transfer tube 20 with inner groove)
In the present embodiment, the inner diameter (ID) of the main pipe 21 is controlled to be not less than 2.4 mm and not more than 6.8 mm. The reason for this is as follows.

  FIG. 3 shows the relationship between the inner diameter and the wall thickness that provides a pressure resistance equivalent to that of an internally grooved tube made of phosphorous-deoxidized copper having an outer diameter of 7 mm and a bottom wall thickness of 0.45 mm (inner diameter 6.1 mm).

  Please refer to FIG. If the inner diameter is increased, the wall thickness must be increased in terms of pressure resistance, which increases the weight. If the inner diameter exceeds 6.8 mm, the bottom wall thickness exceeds 0.5 mm. When the bottom wall thickness is increased, the workability of the inner surface groove is deteriorated, the degree of freedom of the fin shape is lost and not only the cost is increased, but also the outer surface quality of the heat transfer tube is deteriorated. Therefore, the upper limit of the inner diameter (ID) is preferably 6.8 mm.

  FIG. 4 shows a cross section of a schematic configuration of an internally grooved tube processing apparatus for explaining a method of processing an internally grooved heat transfer tube.

  In the internally grooved tube processing apparatus, the grooved plug 4 and the floating plug 2 are connected via a connecting rod 3 in a freely rotating state and inserted into the tube of the copper tube 1. By pulling out the copper tube 1, the floating plug 2 is held at the position of the drawing die 10, thereby fixing the position of the grooved plug 4 in the drawing direction. A ball 6 or a roll as a pressing means revolves around the outer peripheral surface of the copper tube 1 of the fixed portion, thereby forming a fin (groove 8) corresponding to the shape of the groove 5 of the grooved plug 4 in the pipe. The inner surface grooved tube 9 having 8 is further drawn out, and the inner surface grooved tube 9 is reduced in diameter by the die 7 to produce the inner surface grooved heat transfer tube 20.

  In FIG. 4, even if the copper tube 1 is pressed against the groove of the plug 4 with the ball 6, the bottom wall thickness is large, so that the force is not easily transmitted in the radial direction and the axial direction extends. For this reason, the bottom wall thickness of the inner grooved tube 9 becomes smaller than a desired value, and therefore the thickness of the copper tube 1 must be increased excessively. At this time, since the amount of diameter reduction by the ball 6 increases, the outer surface quality of the heat transfer tube deteriorates.

(Regarding the lower limit of the ID of the heat transfer tube 20 with the inner groove)
If the inner diameter is reduced, the pressure loss increases with the number of passes as it is, and the COP is lowered. Therefore, it is necessary to increase the number of passes of the heat exchanger. However, if the number of passes is too large, the shunt cost increases. Note that the number of paths is the number of paths through which the refrigerant flows. If all of the heat transfer tubes are connected continuously, one path is used. If there are two paths connected continuously, two paths are used. That's it. Consider, for example, a case where the heat transfer area is the same and the pressure loss is the same even if the diameter is reduced, compared to a 7 mm outer diameter of a general heat transfer tube for a CO ₂ refrigerant heat pump water heater.

FIG. 5A is an external view of an evaporator flow divider of a CO ₂ refrigerant heat pump water heater that is actually used, and FIG. 5B is a cross-sectional view of a flow dividing portion of a flow divider having an outer diameter of a heat transfer tube of 7 mm. (B) of FIG. 5B shows the assumption cross section of the flow shunt whose outer diameter of a heat exchanger tube is 3 mm. Moreover, FIG. 5C shows the outline at the time of inserting the copper pipe before and after fractionation into the fractionator.

  In FIG. 5A, a heat transfer tube is inserted in the inner circle, and the refrigerant before diversion flows into the flow divider from below, and is divided into a plurality of paths at the diversion portion and goes out from above. The cross-sectional view of the flow dividing portion having an outer diameter of 3 mm shown in FIG. 5B (b) is drawn with the dimensional spacing of the flow dividing portions kept substantially the same in consideration of brazing reliability. In the vicinity of the diameter of 3 mm, about 10 passes is the limit for the same size as the current shunt. If the path is longer than this, the size of the shunt is increased or the number of shunts is increased, both of which lead to an increase in cost and space. Therefore, the pressure loss when the small-diameter tube heat exchanger is configured with a realistic number of passes (10 passes or less) will be calculated and evaluated for smooth tubes hereinafter.

  The pressure loss ΔP in the heat transfer tube installed horizontally is expressed as the sum of the acceleration loss ΔPa and the friction loss ΔPF.

  The acceleration loss ΔPa is expressed by the following equation (2).

Here, G is the refrigerant flux per pass (kg / (m ² s)), ρL is the refrigerant liquid density (kg / m ³ ), ρv is the refrigerant vapor density (kg / m ³ ), and xe is The degree of dryness at the exit.

  In order to obtain the friction loss ΔPF, first, a two-phase flow friction loss gradient (ΔP / Δz) F is obtained. (ΔP / Δz) F is obtained as a ratio to the friction loss gradient (ΔP / Δz) L, (ΔP / Δz) V when only the liquid phase component or the gas phase component flows as a single-phase flow (formula (3 )). In addition, the friction loss gradient of the single-phase flow can be obtained from Equation (4).

  Here, ID in the equation (4) is the inner diameter (mm) of the tube, v is the flow velocity of the single phase (liquid or steam), f is the coefficient of tube friction, and for example, the equation (5) of Blasius is used.

Φ _L ² and Φ _{V in} Equation (3) are two-phase flow doubling coefficients, and are expressed by the following Chisholm-Laird Equations (6) and (7).

  Here, X is a Lockhart-Martinelli parameter and is expressed by the following Colburn equation (8).

  Here, x is the dryness of the two-phase refrigerant, and the average dryness of the inlet / outlet is used. Further, μL is the viscosity (Pas) of the liquid refrigerant, and μV is the viscosity (Pas) of the vapor refrigerant.

  When the total heat transfer area is the same and the number of passes is changed from 2 passes to 10 passes, the heat transfer tube length per pass is obtained by the following equation (9).

  Here, ID is the inner diameter (mm), L is the heat transfer tube length (mm) per pass, and N is the number of passes.

  The heat exchanger composed of a heat transfer tube with an inner diameter of 6.1 mm used as a comparison object has a length of one heat transfer tube of 800 mm, two rows and 30 stages, and a heat transfer tube with an outer diameter of about 3 mm per pass. The length is calculated from equation (9).

  FIG. 6 shows a schematic diagram of a two-row 30-stage two-pass heat exchanger.

  A large number of fin members are provided vertically, and the heat transfer tube has a shape penetrating them. Further, the heat transfer tube is repeatedly deformed into a U-shape and is stored in the fin material on the same plane, and the number of the planes is called a row. In FIG. 6, it has a two-row configuration in front and back. The number of stages indicates the number of heat transfer tubes arranged in the vertical direction, and is the number of the U-shaped portions + 1. In FIG. 6, there are 30 stages. In the schematic diagram of FIG. 6, the two paths are divided into a front row and a back row, and there are two paths through which the refrigerant flows. This example is given as an example for calculating the length of the heat transfer tube, and the configuration of the path, the number of columns, and the number of stages can be changed as appropriate. Further, the friction loss ΔPF can be calculated by substituting for z in the equation (3).

  FIG. 7 shows the result of calculating the pressure loss ratio for an outer diameter of 7 mm (inner diameter of 6.1 mm) × 2 passes when the inner diameter is used as a parameter and the number of passes is 10.

  If the inner diameter is smaller than 2.4 mm, the pressure loss is larger than the inner diameter of 6.1 mm. Therefore, the inner diameter is preferably 2.4 mm or more.

(About the upper limit of fin height [HF])
Groove depth, i.e., when the fin height HF to form a fin exceeding 0.25 mm, it wider base width W ₁ of the fins, so the result becomes narrower groove bottom width W ₃ as, the residence of the refrigerating machine oil And heat transfer rate is reduced. Therefore, HF is preferably 0.25 mm or less.

(About the lower limit of fin height [HF])
FIG. 8 shows a fin height and a smooth inner surface area tube for a heat transfer tube having an inner diameter of 6.8 mm and an inner diameter of 2.4 mm when the ratio (W ₃ / ID) of the groove bottom width to the inner diameter is constant at 0.04. The calculation result of ratio is shown.

  When the fin height is 0.25 mm or less, there is a cost merit that the weight is reduced as the fin height is reduced, but the surface area is also reduced, so that the heat transfer coefficient is gradually lowered. Referring to FIG. 8, HF needs to be 0.1 mm or more in order to secure an internal surface area of 1.5 times or more of the smooth tube ratio in the cross-sectional area.

  In general, in a heat exchanger using a refrigerant that is not affected by refrigerating machine oil, an increase in internal surface area is proportional to heat transfer performance. Here, the refrigerant that is not affected by the refrigeration oil means that the compatibility between the refrigerant and the refrigeration oil is high, and is used in a combination of an alternative chlorofluorocarbon R410A used in an air conditioner or the like and a synthetic oil such as an ether type. Has been.

  Moreover, in the heat exchanger by a heat exchanger tube and a fin material, the ratio of heat resistance is generally said to be the heat exchanger tube inner side: heat exchanger outer side = 3: 7. If the performance of the heat transfer tube is improved by 1.5 times that of the smooth tube, the heat resistance in the heat transfer tube is reduced from 3 to 2, and the heat resistance of the entire heat exchanger can be reduced by 10%. The increase in the weight of the inner grooved tube relative to the smooth tube is approximately 10%, although it depends on the fin shape. Even if the weight is increased by changing from the smooth tube to the inner grooved tube, the thermal resistance can be reduced by 10% or more. The heat exchange length can be shortened by 10% to keep the weight equal, and the size can be reduced as the size is shortened.

(About the lower limit of W ₃ / ID)
Normally, mixing of refrigeration oil into the CO ₂ refrigerant cannot be completely prevented unless an oil separator is provided, and the amount thereof is at least 0.1% by mass or more generally referred to in room air conditioners. Further, when the amount of remaining hot water in the tank is small, the amount of oil flowing out together with the refrigerant increases because the refrigerant circulation amount is increased by increasing the compressor rotation speed.

  FIG. 9 shows an example of the result of measuring the rotation speed ratio of the compressor and the amount of refrigeration oil mixed in the refrigerant.

  The compressor rotation speed ratio on the horizontal axis in FIG. 9 is a ratio in which steady operation is set to 1, and in actual operation, it is necessary to boil hot water during the nighttime power hours. Drive at a higher speed than normal. Considering + 30% as the fluctuation range, it becomes 0.5% by mass or more.

When the amount of refrigerating machine oil mixed into the CO ₂ refrigerant is 0.1% by mass or more, if the ratio W ₃ / ID of the groove bottom width to the inner diameter is less than 0.04, the refrigerating machine oil stays, and the surface area expansion effect by the fins is utilized. The heat transfer rate decreases. In addition, an internally grooved tube having a small W ₃ / ID has a large number of fins or a high fin height, so that the heat transfer tube weight also increases.

(About the upper limit of W ₃ / ID)
On the other hand, as W ₃ / ID increases, the heat transfer tube weight decreases, and the effective heat transfer area for the oil film that greatly affects the heat transfer coefficient increases, but the inner surface area decreases.

FIG. 10 shows the calculation result of W ₃ / ID and the inner surface area smooth tube ratio with the same fin shape for the case where the inner diameter and the fin height are 6.8 mm, 0.25 mm, 2.4 mm, and 0.1 mm. Indicates.

  Referring to FIG. 10, in order to secure an inner surface area of 1.5 times or more of the smooth tube ratio in the cross-sectional area, it is preferable to set it to 0.1 mm or less.

(Effect of embodiment)
The inner grooved heat transfer 20 according to the present embodiment is carbon dioxide having a low heat flux (q) polyalkylene glycol-based refrigeration oil content of 0.1% by mass or more like a CO ₂ refrigerant heat pump water heater. When a refrigerant is used, the groove bottom width is optimized so that the groove depth is sufficiently secured and the stagnation of the refrigerating machine oil can be suppressed. Therefore, a high-performance internally grooved heat transfer tube 20 can be provided.

  Further, since the optimization of the groove bottom width in the internally grooved heat transfer tube having the same fin height is determined by the number of fins and the thickness of the fin, it is possible to prevent the heat transfer tube from becoming heavy.

Further, by increasing the inner diameter to 2.4 mm or more, an increase in pressure loss is suppressed, and by setting the inner diameter to 6.8 mm or less, it is possible to prevent an increase in weight with a thick wall corresponding to a high-pressure CO ₂ refrigerant.

  Examples will be described below. First, the heat transfer coefficient of the heat transfer tube will be described.

  FIG. 11 shows an outline of the measuring apparatus.

  The pressure and temperature before the expansion valve were adjusted with Gas Cooler 1 and Gas Cooler 2, and the refrigerant was distributed to the test section and bypass circuit to control the refrigerant flow rate. The refrigerant flow rate is determined by the expansion valve and the dryness adjusting heat exchanger Heat Ex. , Measured with a Coriolis mass flow meter with an accuracy of ± 0.1%. Further, before the expansion valve, a circuit passing through the sampling cylinder and a bypass circuit were installed in parallel, and the refrigerant flowing in a steady state was sampled to measure the refrigerator oil content (oil concentration). The saturation temperature at the exit of the test section was adjusted by Evapolator 1 and at the exit of the test section, when the refrigerant was a gas-liquid two-phase flow, it was adjusted to take superheat by Evaporator 2.

  FIG. 12 shows an outline of the heat transfer tube measuring section (test section).

  The test section is a double pipe installed horizontally, and a test tube having a heat exchange length of 1 m is used. The inner diameter of the double pipe outer tube is 13 mm, and heat is exchanged by flowing cold / hot water at a constant flow rate in an opposite direction to the annular part. The amount of heat exchange was determined from the inlet temperature Twi, the outlet temperature and Two, and the volumetric flow rate V of the water flowing through the double-tube annular portion. The volumetric flow rate was measured with a volumetric flowmeter composed of a pair of non-circular gears with an accuracy of ± 0.2%. Further, by measuring the temperature of water at the inlet of the flow meter, the density ρw of the water passing through the flow meter was obtained, and the mass flow rate was calculated. Moreover, the temperature of water was measured using the resistance temperature sensor of (PHI) 6.4. The formula for obtaining the heat exchange amount is as shown in formula (10).

  Here, h is the enthalpy (kJ / kg) determined from the temperature and pressure of water.

  On the other hand, the temperature of the refrigerant was measured using a resistance thermometer of φ2, and the pressure was measured using a pressure transducer with a precision of ± 0.08% (BSL notation) of full scale. The heat transfer coefficient in the tube of the heat transfer tube was calculated from the heat transfer rate and the heat transfer coefficient outside the tube. The calculation formula is shown below. The heat transfer rate is calculated using equation (11), the external heat transfer coefficient is calculated using equation (12), and the in-tube heat transfer coefficient is calculated using equation (13).

  Here, ΔTL is the logarithm average temperature difference (K) in the measurement section, and L is the measurement length (m).

  Here, λw is the thermal conductivity (W / (mK)) of the water in the measurement section, De is the equivalent diameter of the water flow path, and Di is the inner diameter (m) of the double outer pipe (water pipe). , Re is the Reynolds number of water in the measurement section, and Pr is the Prandtl number.

The thermophysical values of CO ₂ and water were calculated by PROPATH.

  First, the result of examining the influence of heat flux (q) is shown. Table 1 shows the measurement results of Example A and Comparative Examples A to E.

Example A, Comparative Example A, and Comparative Example B are the same heat transfer tubes. However, as a measurement condition, Comparative Example A has a heat flux (q) of 20 kW / m ² and Comparative Example B has 10 kW / m ² . Example A is the result of measuring the heat transfer coefficient at 5 kW / m ² (in Table 1, W ₃ is the groove bottom width). Similarly, Comparative Examples C to E are the results of measuring the same heat transfer tube while changing the heat flux (q). Other test conditions are a refrigerant saturation temperature of 8 ° C., a refrigerant mass flux of 400 kg / (m ² s), a dryness of 0.35, and an oil concentration of 0.1%.

  The dryness is the mass ratio of the vapor component of the refrigerant that flows in a two-phase flow in which liquid and vapor are mixed. In the evaporator, a refrigerant with a low dryness of about 0.2 flows into the evaporator, absorbs heat from the outside air and gradually evaporates, and becomes a refrigerant with a dryness of 1 or more, which is only vapor, and the evaporator. And is sucked into the compressor.

When the heat flux (q) is 10 kW / m ² or more, Comparative Example D and Comparative Example E with a larger number of fins and a narrow groove bottom width have higher heat transfer rates than Comparative Example A and Comparative Example B. When (q) is 5 kW / m ² , the heat transfer coefficient of Example A having a smaller surface area but wider groove bottom width is higher than that of Comparative Example C. When the heat flux (q) is 20 kW / m ² , the heat transfer coefficient of the inner surface grooved tube of Comparative Example E, which is 23% larger than that of Comparative Example B, is 23% larger. This can be considered as a result of the promotion of nucleate boiling. On the other hand, when the heat flux (q) is 5 kW / m ² , the heat transfer coefficient of Example A is larger than that of Comparative Example C because the heat flux (q) is smaller and the heat transfer due to nucleate boiling. It can be considered that the effect of the oil film formed in Comparative Example C was large and the effective surface area was decreased on the contrary. As a result, the heat transfer coefficient in the CO ₂ refrigerant is affected by oil even when the oil concentration is 0.1% by mass, and the heat flux (q) is also greatly affected. If the heat flux (q) condition is changed, the heat transfer coefficient is changed. It can be seen that the groove shape of the inner grooved tube having a high height is completely different.

Next, the performance of the internally grooved tube shown in Table 2 was measured at a heat flux (q) of 5 kW / m ² (in Table 2, W ₃ is the groove bottom width).

Here, Comparative Example 1 is a smooth tube. The inner grooved pipes of Comparative Example 2 and Examples 1 to 3 have the same fin shape such as the thickness of the fin. The test conditions are a refrigerant saturation temperature of 8 ° C., a refrigerant mass flux of 360 kg / (m ² s), a dryness of 0.35 and 0.75, and an oil concentration of 3%.

  FIG. 13 shows the test results as the relationship between the number of fins and the heat transfer coefficient smooth tube ratio.

  The heat transfer coefficient averages the results of dryness of 0.35 and 0.75. Similar to the results for the oil concentration of 0.1% in Table 1, the heat transfer coefficient increased despite the surface area decreasing as the number of fins decreased.

FIG. 14 shows the relationship between W ₃ / ID and the heat transfer coefficient smooth tube ratio / single-height smooth tube ratio, where the heat transfer coefficient smooth tube ratio is the smooth tube ratio of weight (unit weight) per unit length. The value divided by is on the vertical axis.

This is an index in which the effect of improving the heat transfer coefficient by the internally grooved tube is compared with the weight ratio directly linked to the cost. As the W ₃ / ID increases, the “heat transfer coefficient smooth tube ratio / single weight” Although the “smooth tube ratio” is large, it can be seen that the effect is particularly great when W ₃ / ID is 0.035 or more.

Then, to manufacture a inner grooved tube shown in Table 3 Effect of fin height, it shows the results of investigation (Table W ₃ 3 Among is groove bottom width). The test conditions are: heat flux (q) 5 kW / m ² , refrigerant saturation temperature 8 ° C., refrigerant mass flux 200 kg / (m ² s), dryness 0.35, 0.75, oil concentration 2 %. The heat transfer coefficient averages the results of dryness of 0.35 and 0.75.

W ₃ / ID is 0.04 or more for all, and comparing Example 4 and Example 5, the higher the fin height, the higher the heat transfer coefficient and the ratio of heat transfer coefficient and single weight. Yes. However, in Comparative Example 3, compared with Example 5, the heat transfer rate decreases despite the fin height being higher, and the ratio of heat transfer rate and single weight is compared with Example 4. Is also getting smaller. Therefore, the fin height is desirably 0.25 or less.

  As described above, as shown in the embodiments, in the evaporator of the carbon dioxide refrigerant heat pump water heater, the heat exchange tube configured to use the inner surface grooved heat transfer tube is provided, which provides a high performance inner surface grooved heat transfer tube with little weight increase. The heat exchange efficiency of the vessel can be improved.

  While the embodiments and examples of the present invention have been described above, the embodiments and examples described above do not limit the invention according to the claims. It should be noted that not all combinations of features described in the embodiments and examples are necessarily essential to the means for solving the problems of the invention.

DESCRIPTION OF SYMBOLS 1 Copper pipe 2 Floating plug 3 Connecting rod 4 Grooved plug 5 Groove 6 Ball 7 Drawing die 8 Fin 9 Inside grooved tube 10 Drawing die 11 Compressor 12 Gas cooler 13 Expansion valve 14 Evaporator 20 Inside grooved heat transfer tube 21 Main tube 22 Groove 23 Fin

Claims (3)

Carbon dioxide with a content of polyalkylene glycol refrigerating machine oil of 0.1% by mass or more is used as a refrigerant and used in an evaporator provided in a heat pump water heater used with a heat flux (q) of less than 10 kW / m ² An internally grooved heat transfer tube,
A main pipe having a plurality of spiral grooves provided in parallel to each other on the inner peripheral surface, and a plurality of fins located between the plurality of grooves;
The inner diameter (ID) of the main pipe is 2.4 mm or more and 6.8 mm or less,
The height (HF) of the fin is 0.1 mm or more and 0.25 mm or less,
An internally grooved heat transfer tube that satisfies 0.04 ≦ W ₃ /ID≦0.1, where W ₃ is the bottom width of the plurality of grooves.   The internally grooved heat transfer tube according to claim 1, wherein an angle formed by the groove and the fin is 15 ° or more and 25 ° or less.   The internally grooved heat transfer tube according to claim 2, wherein the plurality of fins is 30 or more and 50 or less.

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