JP5858478B2 - Parallel flow type heat exchanger and air conditioner equipped with the same - Google Patents

Description

  The present invention relates to a side flow parallel flow heat exchanger and an air conditioner equipped with the heat exchanger.

  A parallel flow type heat in which a plurality of flat tubes are arranged between a plurality of header pipes so that a plurality of refrigerant passages in the flat tubes communicate with the inside of the header pipe, and fins such as corrugated fins are arranged between the flat tubes. Exchangers are widely used in outdoor units of car air conditioners and building air conditioners.

  An example of the structure of a parallel flow heat exchanger is shown in FIG. In FIG. 1, the upper side of the paper is the upper side of the heat exchanger, and the lower side of the paper is the lower side of the heat exchanger. The parallel flow type heat exchanger 1 is a side flow type, and includes two vertical header pipes 2 and 3 and a plurality of horizontal flat tubes 4 disposed therebetween. The header pipes 2 and 3 are arranged in parallel in the horizontal direction at intervals, and the flat tubes 4 are arranged at a predetermined pitch in the vertical direction. Since the heat exchanger 1 is installed at various angles according to design requirements at the stage of actually mounting on equipment, the “vertical direction” and “horizontal direction” in this specification should not be strictly interpreted. It should be understood as a mere measure of direction.

  The flat tube 4 is an elongated molded product obtained by extruding a metal, and as shown in FIG. 2, a refrigerant passage 5 through which a refrigerant flows is formed. Since the flat tube 4 is disposed so that the extrusion direction, which is the longitudinal direction, is horizontal, the refrigerant flow direction of the refrigerant passage 5 is also horizontal. A plurality of refrigerant passages 4 having the same cross-sectional shape and the same cross-sectional area are arranged in the left-right direction in FIG. 2, so that the vertical cross-section of the flat tube 4 has a harmonica shape. Each refrigerant passage 5 communicates with the inside of the header pipes 2 and 3.

  Fins 6 are attached to the flat surface of the flat tube 4. Here, corrugated fins are used as the fins 6, but plate fins may be used. Of the fins 6 arranged vertically, the side plate 7 is disposed outside the uppermost and lowermost fins.

  The header pipes 2 and 3, the flat tubes 4, the fins 6, and the side plates 7 are all made of a metal having good heat conduction such as aluminum, the flat tubes 4 are for the header pipes 2 and 3, and the fins 6 are for the flat tubes 4. The side plate 7 is fixed to the fin 6 by brazing or welding.

  The inside of the header pipe 2 is partitioned into three sections S1, S2, and S3 by two partition plates P1 and P2. The partition plates P1 and P2 divide a plurality of flat tubes 4 into three flat tube groups, and a plurality of flat tubes 4 are connected to each of the sections S1, S2, and S3.

  The inside of the header pipe 3 is partitioned into two sections S4 and S5 by a single partition plate P3. The partition plate P3 divides a plurality of flat tubes 4 into two flat tube groups, and a plurality of flat tubes 4 are connected to each of the sections S4 and S5.

  A refrigerant inlet / outlet pipe 8 is connected to the section S1. A refrigerant inlet / outlet pipe 9 is connected to the section S3.

  The function of the heat exchanger 1 is as follows. When the heat exchanger 1 is used as a condenser, the refrigerant is supplied to the section S1 through the refrigerant inlet / outlet pipe 8. The refrigerant that has entered the compartment S1 travels to the compartment S4 through a plurality of flat tubes 4 that connect the compartment S1 and the compartment S4. The flat tube group including the plurality of flat tubes 4 constitutes the refrigerant path A. The refrigerant path A is symbolized by a block arrow. Other refrigerant paths are also symbolized by block arrows.

  The refrigerant that has entered the compartment S4 turns back there, and travels through the plurality of flat tubes 4 connecting the compartment S4 and the compartment S2 to the compartment S2. The flat tube group including the plurality of flat tubes 4 constitutes the refrigerant path B.

  The refrigerant that has entered the section S2 is turned back there, and travels to the section S5 through a plurality of flat tubes 4 that connect the sections S2 and S5. The flat tube group including the plurality of flat tubes 4 constitutes the refrigerant path C.

  The refrigerant that has entered the section S5 is turned back there, and travels to the section S3 through a plurality of flat tubes 4 that connect the sections S5 and S3. The flat tube group including the plurality of flat tubes 4 constitutes the refrigerant path D. The refrigerant entering the section S3 flows out from the refrigerant inlet / outlet pipe 9.

  In this specification, the section from the refrigerant inlet / outlet pipe 8 or 9 to the first turn or between the turn and the next turn is referred to as “one turn”. The refrigerant paths A, B, C, and D are all one-turn refrigerant paths.

  When the heat exchanger 1 is used as an evaporator, the refrigerant is supplied to the section S3 through the refrigerant inlet / outlet pipe 9. The refrigerant flow thereafter follows the refrigerant path when the heat exchanger 1 is used as a condenser. That is, the refrigerant enters the section S <b> 1 through the refrigerant path D → refrigerant path C → refrigerant path B → refrigerant path A and flows out of the refrigerant inlet / outlet pipe 8.

  In parallel flow type heat exchangers, various ideas are devised in order to improve performance. Examples thereof can be seen in Patent Documents 1 to 3.

  In the parallel flow heat exchanger described in Patent Document 1, a fluid diameter of 0.015 inch (about 0.38 mm) to 0.07 is provided inside a plurality of flat tubes connecting two header pipes. A plurality of refrigerant passages in the range of 1 inch (about 1.78 millimeters) are formed in parallel, and the cross-sectional profile of the refrigerant passages is at least two relatively straight portions where they meet and where they meet. It is supposed to have one recessed part. With this configuration, the air-side front area that is blocked by the flat tube is small, and the air-side heat transfer surface can be increased without increasing the air-side pressure drop.

  In the parallel flow type heat exchanger described in Patent Document 2, by setting the height of the refrigerant passage in the flat tube from 0.35 millimeters to 0.8 millimeters, the heat radiation reduction due to the ventilation resistance and the pipe pressure loss are set. The heat dissipation performance is improved by reducing the sum of the decrease in heat dissipation performance due to.

  In the parallel flow type heat exchanger described in Patent Document 3, the shunt parameter γ, which is the ratio of the resistance parameter β of the flat tube to the resistance parameter α of the header pipe on the refrigerant inlet side, is set to 0.5 or more, and the header pipe Blocks the refrigerant from intensively flowing into the flat tube connected to the highest pressure part of the refrigerant inlet, uniforming the pressure applied to each flat tube to obtain a good diversion state and good heat exchange performance Is to be demonstrated.

Japanese Patent Laid-Open No. 5-87752 JP 2001-165532 A Japanese Patent Laid-Open No. 2000-111274

  When a parallel flow type heat exchanger is used as an evaporator, considering a refrigerant flowing in the refrigerant path, a liquid refrigerant flows in a certain flat tube, and a large amount of gaseous refrigerant flows in another flat tube. It is desirable that the state of “uneven flow” does not occur. An object of the present invention is to provide a side flow parallel flow type heat exchanger that is optimally designed with respect to the number of flat tubes constituting a refrigerant path, from the viewpoint of preventing drift. In particular, the object is to optimize the number of flat tubes in the refrigerant path having a large proportion of gaseous refrigerant.

The parallel flow heat exchanger according to the present invention is characterized by the following configuration. That is, two vertical header pipes and a plurality of horizontal flat tubes connecting the header pipes are provided, and the plurality of horizontal flat tubes are further grouped in groups, Constitutes a one-turn refrigerant path for flowing refrigerant from one of the two vertical header pipes to the other, and the upper limit of the number of the flat tubes constituting the one-turn refrigerant path is as follows: Determined by the numerical value ± 2 obtained by Equation A:
When using the parallel flow type heat exchanger for an outdoor unit of an air conditioner,
n <3.0 × 10 ⁻⁴ × Q + 8.0 (A)
When using the parallel flow type heat exchanger for an indoor unit of an air conditioner,
n <4.2 × 10 ⁻⁴ × Q + 7.9 (A)
However, n is the number of flat tubes constituting a one-turn refrigerant path, Q is a rated capacity, and W is a unit. Q is a rated heating capacity for an outdoor unit and a rated cooling capacity for an indoor unit.

When the parallel flow type heat exchanger having the above configuration is used in an outdoor unit of an air conditioner, the lower limit of the number of the flat tubes constituting the one-turn refrigerant path is preferably determined by the following formula B:
n> (αQ + β) × {(1.4 × 10 ⁻¹⁶ ) × L / (d × A ′ ² )} ^0.5 (B)
Where α = 0.0161, β = 8.86, d is a hydraulic diameter in units of m, A ′ is a cross-sectional area of the refrigerant passage of one flat tube, and m ² is in units. To do.

When the parallel flow type heat exchanger having the above configuration is used in an indoor unit of an air conditioner, the lower limit of the number of the flat tubes constituting the one-turn refrigerant path is preferably determined by the following formula B:
n> (αQ + β) × {(1.4 × 10 ⁻¹⁶ ) × L / (d × A ′ ² )} ^0.5 (B)
Where α = 0.0228, β = 6.62, d is a hydraulic diameter in m, and A ′ is a cross-sectional area of the refrigerant passage of one flat tube, and m ² is in units. To do.

  Further, the present invention is an air conditioner in which the parallel flow heat exchanger having the above-described configuration is mounted on an outdoor unit or an indoor unit.

  According to the present invention, it is possible to obtain a side flow parallel flow type heat exchanger that does not cause a drift according to the circulation amount of the refrigerant.

It is a schematic block diagram of the parallel flow type heat exchanger of a side flow system. It is sectional drawing along the II-II line of FIG. It is a table | surface of the specification of the test product of a flat tube. It is a table | surface which shows the relationship between a refrigerant | coolant circulation amount and the number of flat tubes which do not produce a drift. It is a graph which shows the relationship between a refrigerant | coolant circulation amount and the number of flat tubes. It is a graph which shows the relationship between the air_conditioning | cooling capability and refrigerant | coolant circulation amount. It is a graph which shows the relationship between heating capability and a refrigerant | coolant circulation amount. It is a graph of the flat tube number optimal range regarding an air conditioner outdoor unit. It is a graph of the flat tube number optimal range regarding an air conditioner indoor unit. It is a graph which shows the relationship between a refrigerant | coolant circulation amount and a suction pressure. It is a graph which shows the relationship between a refrigerant | coolant circulation amount and the number of flat tubes. It is a graph which shows the relationship between the number of the flat tubes in a heat exchanger for outdoor units, and a heating rated capacity. It is a graph which shows the relationship between the number of the flat tubes in a heat exchanger for indoor units, and a cooling rated capacity. It is a schematic block diagram of the air conditioner carrying the parallel flow type heat exchanger which concerns on this invention, and shows the state at the time of heating operation. It is a schematic block diagram of the air conditioner carrying the parallel flow type heat exchanger which concerns on this invention, and shows the state at the time of air_conditionaing | cooling operation.

  The parallel flow type heat exchanger 1 of the side flow type shown in FIG. 1, in which the number of flat tubes 4 constituting the refrigerant path is set by the method described below, is a parallel flow type heat according to the present invention. It shall be an exchanger. However, the number of refrigerant paths is not limited to four. It may be more or less.

First, the upper limit of the number of flat tubes 4 constituting the one-turn refrigerant path is obtained. The upper limit value is obtained from the following formula A.
When using the parallel flow type heat exchanger for an outdoor unit of an air conditioner,
n <3.0 × 10 ⁻⁴ × Q + 8.0 (A)
When using the parallel flow type heat exchanger for an indoor unit of an air conditioner,
n <4.2 × 10 ⁻⁴ × Q + 7.9 (A)
However, n is the number of flat tubes constituting a one-turn refrigerant path, Q is a rated capacity, and W is a unit.

Formula A was derived from the test. The table of FIG. 3 shows the specifications of the flat tube examined in the test. The test product a has a width of 16.2 mm, a thickness of 1.9 mm, and a refrigerant passage cross-sectional area of 13 mm ² . The test product b has a width of 13.9 mm, a thickness of 1.9 mm, and a refrigerant passage cross-sectional area of 11 mm ² . The test product c has a width of 16.2 mm, a thickness of 1.6 mm, and a refrigerant passage cross-sectional area of 11 mm ² . The test product d has a width of 19.2 mm, a thickness of 1.9 mm, and a refrigerant passage cross-sectional area of 14 mm ² .

  The test was conducted as follows. The refrigerant is circulated through various numbers of flat tubes, and whether or not a drift has occurred is visually confirmed by thermography. Using the four types of test products shown in FIG. 3, the refrigerant was circulated by changing the circulation amount for each test product, and no drift was observed in the refrigerant circulation amount (this state is referred to as “no drift current” in this specification). The table of FIG. 4 summarizes the maximum number of flat tubes.

  In the table of FIG. 4, test 1 used test product a. When the refrigerant circulation rate is 27.3 kg / h, the maximum number of no drift current is 8. When the refrigerant circulation rate is 42.5 kg / h, the maximum number of no drift currents is 9, and when the refrigerant circulation rate is 64.3 kg / h. The maximum number of no-flow currents was 10, and the maximum number of no-flow currents was 10 when the refrigerant circulation rate was 63.2 kg / h.

  In test 2, test product b was used. When the refrigerant circulation rate was 20.9 kg / h, the maximum no-flow current was 9 and when the refrigerant circulation rate was 22.1 kg / h, the maximum no-flow current was 8.

  In test 3, test product c was used. When the refrigerant circulation rate is 59.2 kg / h, the maximum number of non-biased flows is 10, when the refrigerant circulation rate is 48.8 kg / h, the maximum number of no-biased flows is 9, and when the refrigerant circulation rate is 26.4 kg / h The maximum number of non-biased currents was 8.

  In test 4, test product b was used. When the refrigerant circulation rate was 54.8 kg / h, the maximum no-flow current was 8. When the refrigerant circulation rate was 89.2 kg / h, the maximum no-flow current was 8.

  In test 5, test product d was used. When the refrigerant circulation rate is 26.6 kg / h, the maximum number of non-biased flows is 6, when the refrigerant circulation rate is 44.3 kg / h, the maximum number of no-biased flows is 9, and when the refrigerant circulation rate is 67.3 kg / h The maximum number of non-biased currents was 9.

When the test results of FIG. 4 are plotted, the graph of FIG. 5 is obtained. Draw an approximate line, and from the approximate expression of the straight line approximation,
n = 1.9 × 10 ⁻² m + 7.8 (a)
It will be ± 2.

  The refrigerant circulation amount m (kg / h) is generally set as a value proportional to the rated capacity of the product. The relationship between the refrigerant circulation amount and the capacity is shown in FIGS.

When expressed in numerical formula using the rated heating capacity Q (unit is W),
m = 0.0161Q + 8.86 (b)
It can be expressed.

When expressed by a mathematical formula using the cooling rated capacity Q (unit is W),
m = 0.0228Q + 6.621 (c)
It can be expressed.

The relationship between rated capacity and refrigerant circulation varies somewhat depending on the product. The refrigerant circulation amount is simply calculated from the following formula.
Refrigerant circulation amount m = Compressor rotation speed × Suction pressure density × Compressor volume

  The parallel flow heat exchanger becomes an evaporator during heating operation when used as an outdoor unit heat exchanger for an air conditioner, and during cooling operation when used as a heat exchanger for an indoor unit of an air conditioner. It becomes an evaporator.

Therefore, as shown in FIG. 8, when the parallel flow type heat exchanger is used as an outdoor unit heat exchanger, the flat tube constituting the one-turn refrigerant path is obtained from the above formulas (a) and (b). The upper limit of the number is
n = 3.0 × 10 ⁻⁴ Q + 8.0
It becomes.

As shown in FIG. 9, when the parallel flow heat exchanger is used as an indoor unit heat exchanger, the number of flat tubes constituting one turn of the refrigerant path is calculated from the above formulas (a) and (c). The upper limit is
n = 4.2 × 10 ⁻⁴ Q + 7.9
By using ± 2, it is possible to suppress the drift.

Subsequently, the lower limit of the number of flat tubes 4 constituting each refrigerant path is obtained. The outlet temperature of the heat exchanger is
T _out <0 ℃
Then, the suction pressure is greatly reduced as shown in FIG. That is, the suction pressure rapidly decreases with respect to the refrigerant circulation amount. This is due to frost formation due to the outlet temperature being below 0 ° C.

If the temperature drop due to pressure loss ΔP is T _Dp ,
T _Rin −T _Dp <0 degree. T _Rin is the refrigerant inlet evaporation temperature. The unit of pressure loss ΔP is Pa.

That means
P _Rin −ΔP> P _lim
It becomes. P _Rin is the inlet evaporation pressure, and P _lim is the saturation pressure of the refrigerant at 0 ° C.

here,
ΔP = λ × L / d × ρ × u 2/2
It becomes. λ is the coefficient of friction between the inner wall of the flat tube 4 and the refrigerant, L is the pipe length in m, d is the hydraulic diameter in m, and ρ is the refrigerant density. The unit is kg / m ³ , and u is the flow rate of the refrigerant and is m / s.

The flow velocity u can be obtained from the following equation.
u = M / ρA
M is the amount of refrigerant circulated in units of kg / s, and A is the sum of the refrigerant passage cross-sectional areas of a plurality of flat tubes 4 constituting a one-turn refrigerant path in units of m ² It is.

Therefore,
ΔP = λ / 2ρ × L / dA ² × M ²
It becomes.

Here, if the refrigerant passage cross-sectional area of one flat tube 4 is A ′,
A = nA '
It becomes. n is the number of the flat tubes 4 constituting the one-turn refrigerant path.

here,
ΔP <P _Rin −P _lim
Λ / 2ρ × L / (dn ² × A ′ ² ) × M ² <P _Rin −P _lim
It becomes.

here,
n ² > M ² × λ / 2ρ × L / dA ′ ² × 1 / (P _Rin −P _lim )
It becomes.

From the above equation: n> M {λ / 2ρ × L / dA ′ ² × 1 / (P _Rin −P _lim )} ^0.5 (d)
It becomes.

The refrigerant circulation amount m (kg / h), which is a unit difference of M, is generally set as a value proportional to the rated capacity of the product. Therefore,
m = αQ + β
It can be expressed.

The relationship between the refrigerant circulation amount and the capacity is shown in FIGS. Expressed as a mathematical expression using the rated heating capacity Q (unit: W), m = 0.0161Q + 8.86
It can be expressed. That is, α = 0.0161 and β = 8.86.

Moreover, when expressed by a mathematical formula using the cooling rated capacity Q (unit is W), m = 0.0228Q + 6.62.
It can be expressed. That is, α = 0.0228 and β = 6.62.
In the case of the outdoor unit heat exchanger, the heating rated capacity may be used, and in the case of the indoor unit heat exchanger, the cooling rated capacity may be used.

The relationship between rated capacity and refrigerant circulation varies somewhat depending on the product. The refrigerant circulation amount is simply calculated from the following formula.
Refrigerant circulation amount m = Compressor rotation speed × Suction pressure density × Compressor volume

Further, the pressure loss is usually suppressed to 200 kPa or less. Therefore,
P _Rin −P _lim <200 × 10 ³

The friction coefficient λ varies depending on the refrigerant circulation amount, the refrigerant pressure, the shape of the flat tube, and the like. Generally, it is about 0.5 to 0.05 in a domestic air conditioner. Moreover, although density (rho) changes with the pressure and dryness of a refrigerant | coolant, in the case of a gaseous refrigerant | coolant, it is 20-70 kg / m < ³ > generally.

From the above, when M is converted to m, n> (αQ + β) × {Π × L / (d × A ′ ² )} ^0.5
Π
1.4 × 10 ⁻¹⁶ <Π <4.8 × 10 ⁻¹⁵
If the lower limit number exceeds the calculation result of the upper limit number according to Formula A, it is desirable to branch in the middle of the inlet or the heat exchanger.
Here, since it is desirable that the pressure loss is low, it is desirable to use the lowest value of Π, that is, 1.4 × 10 ⁻¹⁶ .
Therefore,
n> (αQ + β) × {(1.4 × 10 ⁻¹⁶ ) × L / (d × A ′ ² )} ^0.5 (B)

  From the above, the lower limit of the number of the flat tubes 4 constituting the one-turn refrigerant path can be obtained from Formula B.

FIG. 12 and FIG. 13 are graphs showing an example of the calculation result by Formula B.
FIG. 12 shows the relationship between the number of flat tubes in the outdoor unit heat exchanger and the rated heating capacity, and FIG. 13 shows the relationship between the number of flat tubes in the indoor unit heat exchanger and the rated cooling capacity. In this graph, the number of flat tubes 4 constituting a one-turn refrigerant path is optimized according to the rated capacity, and the lower limit value is shown.

  The parallel flow heat exchanger 1 can be mounted on a separate air conditioner. A separate type air conditioner is composed of an outdoor unit and an indoor unit. The outdoor unit includes a compressor, a four-way valve, an expansion valve, an outdoor heat exchanger, an outdoor fan, and the like. The indoor unit is an indoor heat exchanger, a room Includes an internal blower. The outdoor heat exchanger functions as an evaporator during heating operation and functions as a condenser during cooling operation. The indoor heat exchanger functions as a condenser during heating operation and functions as an evaporator during cooling operation.

  FIG. 14 shows a basic configuration of a separate type air conditioner that uses a heat pump cycle as a refrigeration cycle. The heat pump cycle 101 includes a compressor 102, a four-way valve 103, an outdoor heat exchanger 104, a decompression / expansion device 105, and an indoor heat exchanger 106 connected in a loop. The compressor 102, the four-way valve 103, the heat exchanger 104, and the decompression / expansion device 105 are accommodated in the casing of the outdoor unit, and the heat exchanger 106 is accommodated in the casing of the indoor unit. An outdoor fan 107 is combined with the heat exchanger 104, and an indoor fan 108 is combined with the heat exchanger 106. The blower 107 includes a propeller fan, and the blower 108 includes a cross flow fan.

  The parallel flow type heat exchanger 1 according to the present invention can be used as a component of the heat exchanger 106 of the indoor unit. The heat exchanger 106 is a combination of three heat exchangers 106A, 106B, 106C like a roof that covers the blower 108, and any one of the heat exchangers 106A, 106B, 106C is a parallel flow heat exchanger. 1 can be used.

  The parallel flow type heat exchanger 1 according to the present invention can also be used as the heat exchanger 104 of an outdoor unit.

  FIG. 14 shows a state during heating operation. At this time, the high-temperature and high-pressure refrigerant discharged from the compressor 102 enters the indoor heat exchanger 106 where it dissipates heat and condenses. The refrigerant exiting the heat exchanger 106 enters the outdoor heat exchanger 104 from the decompression / expansion device 105 and expands there, takes heat from the outdoor air, and returns to the compressor 102. The airflow generated by the indoor fan 108 promotes heat dissipation from the heat exchanger 106, and the airflow generated by the outdoor fan 107 accelerates heat absorption of the heat exchanger 104.

  FIG. 15 shows a state during cooling operation or defrosting operation. At this time, the four-way valve 103 is switched so that the refrigerant flow is reversed from that during the heating operation. That is, the high-temperature and high-pressure refrigerant discharged from the compressor 102 enters the outdoor heat exchanger 104, where it dissipates heat and condenses. The refrigerant exiting the heat exchanger 104 enters the heat exchanger 106 on the indoor side from the decompression / expansion device 105 and expands there, takes heat from the indoor air, and returns to the compressor 102. The airflow generated by the outdoor fan 107 promotes heat dissipation from the heat exchanger 104, and the airflow generated by the indoor fan 108 promotes heat absorption of the heat exchanger 106.

  Although the embodiments of the present invention have been described above, the scope of the present invention is not limited to these embodiments, and various modifications can be made without departing from the spirit of the invention.

  The present invention is widely applicable to side flow parallel flow heat exchangers.

1 Heat exchanger 2, 3 Header pipe 4 Flat tube 5 Refrigerant passage 6 Fin 7 Side plate A, B, C, D Refrigerant path

Claims (4)

Two vertical header pipes and a plurality of horizontal flat tubes connecting the header pipes are further provided, and the plurality of horizontal flat tubes are further grouped therein, each group being In the parallel flow type heat exchanger of the side flow type that constitutes a one-turn refrigerant path for flowing the refrigerant from one of the two vertical header pipes to the other,
The parallel flow heat exchanger is used for an outdoor unit of an air conditioner,
The upper limit of the number of the flat tubes constituting the one-turn refrigerant path is determined by a numerical value ± 2 obtained by the following formula A, and the lower limit of the number of the flat tubes constituting the one-turn refrigerant path is parallel flow heat exchangers you characterized in that it is defined by the following formula B:
n <3.0 × 10 ⁻⁴ × Q + 8.0 (A)
n> (αQ + β) × {(1.4 × 10 ⁻¹⁶ ) × L / (d × A ′ ² )} ^0.5 (B)
Where n is the number of flat tubes constituting a one-turn refrigerant path, Q is the rated capacity and is in units of W, α = 0.0161, β = 8.86, and d is the hydraulic diameter. A ′ is the cross-sectional area of the refrigerant passage of one flat tube, and m ² is the unit. Two vertical header pipes and a plurality of horizontal flat tubes connecting the header pipes are further provided, and the plurality of horizontal flat tubes are further grouped therein, each group being In the parallel flow type heat exchanger of the side flow type that constitutes a one-turn refrigerant path for flowing the refrigerant from one of the two vertical header pipes to the other,
The parallel flow heat exchanger is used for an indoor unit of an air conditioner,
The upper limit of the number of the flat tubes constituting the one-turn refrigerant path is determined by a numerical value ± 2 obtained by the following formula A, and the lower limit of the number of the flat tubes constituting the one-turn refrigerant path is parallel flow heat exchangers you characterized in that it is defined by the following formula B:
n <4.2 × 10 ⁻⁴ × Q + 7.9 (A)
n> (αQ + β) × {(1.4 × 10 ⁻¹⁶ ) × L / (d × A ′ ² )} ^0.5 (B)
Where n is the number of flat tubes constituting the one-turn refrigerant path, Q is the rated capacity and is in units of W, α = 0.0228, β = 6.62, d is the hydraulic diameter. A ′ is the cross-sectional area of the refrigerant passage of one flat tube, and m ² is the unit.   An air conditioner comprising the parallel flow heat exchanger according to claim 1 mounted on an outdoor unit.   An air conditioner comprising the parallel flow heat exchanger according to claim 2 mounted in an indoor unit.

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