HFC32 vaporisation inside a Brazed Plate Heat Exchanger (BPHE): Experimental measurements and IR thermography analysis

Accepted Manuscript HFC32 Vaporisation Inside A Brazed Plate Heat Exchanger (Bphe): Experimental Measurements And Ir Thermography Analysis Giovanni A. Longo, Simone Mancin, Giulia Righetti, Claudio Zilio PII:

S0140-7007(15)00115-2

DOI:

10.1016/j.ijrefrig.2015.04.017

Reference:

JIJR 3032

To appear in:

International Journal of Refrigeration

Received Date: 2 March 2015 Revised Date:

20 April 2015

Accepted Date: 21 April 2015

Please cite this article as: Longo, G.A., Mancin, S., Righetti, G., Zilio, C., HFC32 Vaporisation Inside A Brazed Plate Heat Exchanger (Bphe): Experimental Measurements And Ir Thermography Analysis, International Journal of Refrigeration (2015), doi: 10.1016/j.ijrefrig.2015.04.017. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

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HFC32 VAPORISATION INSIDE A BRAZED PLATE HEAT EXCHANGER (BPHE): EXPERIMENTAL MEASUREMENTS AND IR THERMOGRAPHY ANALYSIS Giovanni A. LONGO*, Simone MANCIN, Giulia RIGHETTI, Claudio ZILIO University of Padova, Department of Management and Engineering,

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Str.lla S. Nicola 3, Vicenza, I-36100, ITALY

ABSTRACT

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This paper presents HFC32 average boiling heat transfer coefficients and pressure drops measured inside a small Brazed Plate Heat Exchanger (BPHE): the effects of heat flux, saturation temperature

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(pressure), and outlet conditions are investigated. The experimental tests were carried out at four different saturation temperatures (5, 10, 15, and 20°C) and four different evaporator outlet conditions (vapour quality around 0.80 and 1.00, vapour super-heating around 5 and 10 °C). The average heat transfer coefficients show great sensitivity to heat flux and outlet conditions and weak

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sensitivity to saturation temperature (pressure). The saturated boiling heat transfer coefficients were compared with a new model for refrigerant vaporisation inside BPHE (Longo et al., 2015): the mean absolute percentage deviation between calculated and experimental data is 4.7%. The heat

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transfer and pressure drop measurements are complemented with a IR thermography analysis for a

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better understanding of the vaporisation process inside a BPHE.

*

Corresponding Author ([email protected])

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KEYWORDS: boiling; super-heating; GWP; BPHE; thermography

b

height of the corrugation, m

BPHE

brazed plate heat exchanger

CFC

chlorofluorocarbon

cp

specific heat capacity, J kg-1K-1

CRa

roughness parameter in eq. 13

dh

hydraulic diameter, dh = 2 b, m

f

Fanning friction factor

f.s.

full scale

g

gravity acceleration, m s-2

G

mass flux, G = m/(nchW b), kg m-2s-1

GWP

global warming potential

h

heat transfer coefficient, W m-2 K-1

HC

hydrocarbon

HCFC

hydrochlorofluorocarbon

J k

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HFC

hydrofluorocarbon

specific enthalpy, J kg-1

coverage factor

KE

kinetic energy, J

IR

infra-red

L

flow length of the plate, m

m

mass flow rate, kg s-1

N

number of effective plates

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nominal area of a plate, m2

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A

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NOMENCLATURE

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number of channels

ODP

ozone depleting potential

p

pressure, Pa

p*

reduced pressure

P

corrugation pitch, m

Pr

Prandtl number, Pr = µ cp / λ

q

heat flux, q = Q / S, Wm-2

Q

heat flow rate, W

Ra

arithmetic mean roughness (ISO4271/1), µm

Reeq

equivalent Reynolds number

Rp

roughness (DIN 4762/1), µm

S

nominal heat transfer area, m2

s

plate wall thickness, m

T

temperature, K

U

overall heat transfer coefficient, Wm-2K-1

v

specific volume, m3 kg-1

V

volume, m3

W

width of the plate, m

X

vapour quality, X = (J – JL) /∆JLG

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Greek symbols β

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nch

inclination angle of the corrugation. °

Φ

enlargement factor

λ

thermal conductivity, W m-1K-1

µ

viscosity, kg m-1s-1

ρ

density, kg m-3

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Subscript momentum

boil

boiling

c

manifolds and ports

cb

convective boiling

f

frictional

G

vapour phase

g

gravity

in

inlet

L

liquid phase

LG

liquid vapour phase change

ln

logarithmic

m

average value

nb

nucleate boiling

out

outlet

p

plate

pb

pre-evaporator

r

refrigerant

t

total

sup w

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sat

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a

saturation

super-heating water

wi

water inlet

wo

water outlet

0

reference condition

1.

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INTRODUCTION

Since the early 1980s, National, European, and International regulations had a large impact on the allowed working fluids for refrigeration and air conditioning applications. In the past the Montreal Protocol (United Nations, 1987) determined the phase out of Chlorofluorocarbon (CFCs) and

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Hydrochlorofluorocarbon (HCFCs) refrigerants for their Ozone Depleting Potential (ODP). More recently, the EU F-gas regulation (Regulation (EC) No 842/2006) was aimed to a progressive reduction in the use or to a complete phase out of the Hydrofluorocarbon (HFCs) refrigerants with

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high Global Warming Potential (GWP). Nowadays, the more recent release of the EU F-gas

the market of refrigeration equipment:

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regulation (Regulation (EU) No 517/2014) introduces the following prohibitions on the placing on

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domestic refrigerators and freezers that contain HFCs with GWP ≥ 150 (1 January 2015);

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refrigerators and freezers for commercial use that contain HFCs with GWP ≥ 2500 (1 January 2020) or with GWP ≥ 150 (1 January 2022);

stationary refrigeration equipment that contain HFCs with GWP ≥ 2500 (1 January 2020);

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multipack centralised refrigeration system for commercial use with a rated capacity of 40 kW or

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-

more that contain HFCs with GWP ≥ 150 (1 January 2022); movable room air-conditioning equipment that contain HFCs with GWP ≥ 150 (1 January

-

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2020);

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single split air-conditioning systems containing less than 3 kg of F-gases that contain HFCs with GWP ≥ 750 (1 January 2025).

Therefore most of the currently used HFC refrigerants will be prohibited in the near future by the European Union regulation in different applications. For example HFC134a, showing a GWP of 1430, won’t be used anymore in domestic refrigeration and mobile air-conditioning systems. HFC404A, that exhibits the highest GWP value (3922) among the commonly used refrigerants, will not be applied in commercial refrigeration, and HFC410A, that presents a GWP of 2090, will not be the leading fluid for split air-conditioners, residential heat pumps, and medium size chillers.

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HFC32 exhibits a GWP around 675 and excellent heat transfer and pressure drop performance both in condensation and vaporisation. Therefore, in spite of its mild flammability, it seems to be a very promising low GWP substitute for traditional HFC refrigerants. In the past HFC32 was extensively used as component in refrigerant mixtures, such as HFC410A, whereas its direct use, as pure fluid,

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was rather scarce, probably due to its mild flammability (ANSI / ASHRAE Standard 34, 2013). Daikin has been using HFC32 in split air-conditioning systems since 2012 and, nowadays, HFC32 is currently used in residential and commercial air-conditioners in Japan, China, and India. HFC32

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is also a very promising low GWP refrigerant in medium size chillers and residential heat pumps that implement Brazed Plate Heat Exchangers (BPHEs). The BPHEs, which lead to a reduction of

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the refrigerant charge of one order of magnitude as compared to the traditional tubular heat exchangers, are particularly interesting for reducing the risk of the use of flammable or mildly flammable refrigerants such as HFC32 (Palm, 2007). In fact the first attempt to reduce the risk of flammable refrigerants is to decrease the refrigerant charge.

In the open literature the experimental data on HFC32 two-phase heat transfer inside BPHEs is

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rather scarce. Palmer et al. (2000) measured the average Nusselt number during refrigerant mixture HFC32 / HFC152a (50/50 wt%) vaporisation and condensation inside a BPHE in presence of

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lubricant oil. The performance of this mixture was compared to HC refrigerant (HC290) and HC refrigerant mixture (HC290 / HC600a (70/30 wt.%) ). Mancin et al. (2013) presented HFC32 super-

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heated vapour condensation data inside a BPHE with refrigerant mass flux from 13 to 37 kg m-2 s-1 finding heat transfer coefficients higher than those of HFC410A and HFC407C. Jung et al. (2014) assessed HFC32 as environment friendly substitute for HCFC22 in condensation and vaporisation inside a Plate Heat Exchanger (PHE) for Ocean Thermal Energy Conversion (OTEC), while Lee et al. (2014) investigated the application of the refrigerant mixture HFC32 / HFC152a in a seawater heat pump equipped with PHE. Bella et al. (2014) compared the performance of HFC32 to that of HFC410A in a 70 kW packaged air cooled water chiller which presented a BPHE evaporator. HFC32 was demonstrated to be an effective low GWP substitute for HFC410A in this specific

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application. Longo et al. (2015) investigated HFC32 condensation inside a commercial BPHE and compared its performance to those of traditional HFC refrigerant, and particularly HFC410A. HFC32 exhibits heat transfer coefficients much higher and frictional pressure drop slightly higher than those of HFC410A.

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This paper presents the average heat transfer coefficients and pressure drops measured during HFC32 vaporisation inside a BPHE: the effects of heat flux, saturation temperature (pressure), and outlet conditions are investigated. The heat transfer and pressure drop measurements are

EXPERIMENTAL MEASUREMENTS

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2.

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complemented with the IR thermography analysis carried out during the vaporisation tests.

The experimental facility, shown in Figure 1, consists of a refrigerant loop, a water-glycol loop, and a refrigerated water loop: Table 1 summarises the characteristics of the instrumentation. The evaporator tested is a BPHE consisting of 10 plates, 72 mm in width and 278 mm in length, which

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present a macro-scale herringbone corrugation with an inclination angle of 65° and a corrugation amplitude of 2 mm. Figure 2 and Table 2 give the main geometrical characteristics of the BPHE

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evaporator.

The experimental measurements are reported in terms of average heat transfer coefficients hr and

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frictional pressure drop ∆pf on the refrigerant side. The average heat transfer coefficient hr is computed from the overall heat transfer coefficient U by determining the water side heat transfer coefficient hw.

-1

hr. = (1 / U - s / λp - 1 / hw)

(1)

The overall heat transfer coefficient in the condenser U is equal to the ratio between the heat flow rate Q, the nominal heat transfer area S and the logarithmic mean temperature difference ∆Tln:

ACCEPTED MANUSCRIPT U = Q / (S ∆Tln)

(2)

The heat flow rate Q is derived from a thermal balance on the waterside of the evaporator, the

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reference nominal heat transfer area S is the projected area of the plate. When the evaporator works only in two-phase heat transfer, the logarithmic mean temperature difference is computed with

(3)

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∆Tln=(Twi-Two) / ln[(Twi-Tsat)/(Two-Tsat)]

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reference to the average saturation temperature Tsat on the refrigerant side:

When the evaporator works both in vaporisation and super-heating, according to Dutto et al. (1991) and Fernando et al. (2004), the logarithmic mean temperature difference is computed as the average value between the logarithmic mean temperature difference of the boiling zone and that of the

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super-heating zone weighted on the base of the respective heat flow rates exchanged:

(4)

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∆Tln=Q / [(Qboil / ∆Tln.boil) + (Qsup / ∆Tln.sup)]

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where Qboil and Qsup are the heat flow rate exchanged in the boiling and super-heating zones, ∆Tln.boil and ∆Tln.sup are the logarithmic mean temperature differences in the boiling and super-heating zones. This approach computes the average heat transfer coefficient of the whole evaporator as the average value between the heat transfer coefficient of the boiling zone and that of the super-heating zone weighted on the base of the respective heat transfer area. In this way it is possible to directly compare the heat transfer performance of an evaporator working only in two-phase heat transfer with that of an evaporator working also in vapour super-heating. The water side heat transfer coefficient hw is computed by the following non-dimensional equation:

ACCEPTED MANUSCRIPT hw = 0.277 (λw / dh) Rew0.766 Prw0.333 5 < Prw < 10

(5)

200 < Rew < 1200

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obtained by using a modified Wilson plot technique. The refrigerant vapour quality at the evaporator inlet and outlet Xin and Xout are computed starting from the refrigerant temperature Tpb.in and pressure ppb.in at the inlet of the pre-evaporator (sub-

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cooled liquid condition) considering the heat flow rate exchanged in the pre-evaporator and in the evaporator Qpb and Q and the pressure at the inlet and outlet pin and pout of the evaporator.

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The boiling frictional pressure drop ∆pf is obtained from the total pressure drop measured ∆pt by subtracting the momentum pressure drop ∆pa, the gravity pressure drop ∆pg, and the manifolds and ports pressure drop ∆pc :

(6)

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∆pf = ∆pt - ∆pc - ∆pa - ∆pg

The momentum and gravity pressure drops are estimated by the homogeneous model for two-phase

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flow as follows:

(7)

∆pg = g ρm L

(8)

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∆pa = G2(vG - vL) |∆X|

where G is the refrigerant mass flux, vL and vG are the specific volume of liquid and vapour phase, |∆X| is the absolute value of the vapour quality change between inlet and outlet and ρm is the average two-phase density between inlet and outlet calculated by the homogeneous model at the

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average vapour quality Xm between inlet and outlet. The manifold and port pressure drops are empirically estimated, in accordance with Shah and Focke (1988), as follows

∆pc = 1.5 G2 / (2 ρm)

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(9)

Longo and Gasparella (2007) give an exhaustive description of the experimental set-up, the operating procedures, and data reduction.

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A set of 138 vaporisation tests with refrigerant up-flow and water down-flow was carried out at four different saturation temperatures (5, 10, 15, and 20 °C) and four different evaporator

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outlet conditions (vapour quality around 0.80 and 1.00, vapour super-heating around 5 and 10 °C), whereas the inlet vapour quality ranged between 0.19 and 0.33. Table 3 gives the main operating conditions in the evaporator under experimental tests: refrigerant saturation temperature Tsat and pressure psat, inlet and outlet refrigerant vapour quality Xin and Xout,

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outlet refrigerant super-heating ∆Tsup, mass flux on refrigerant side Gr and water side Gw, and heat flux q. A detailed error analysis performed in accordance with Kline and McClintock (1953) indicates an overall uncertainty within ±12.0% for the refrigerant heat

measurement.

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transfer coefficient measurement and within ±17.3% for the total pressure drop

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Figures 3, 4, and 5 show the average heat transfer coefficients on the refrigerant side against heat flux for three different saturation temperatures (5, 10 and 20 °C) and four different evaporator outlet conditions (vapour quality around 0.80 and 1.00, vapour super-heating around 5 and 10 °C). The heat transfer coefficients show great sensitivity to heat flux and outlet condition and weak sensitivity to saturation temperature (pressure). The saturated boiling heat transfer coefficients with an outlet vapour quality around 0.80 are 6-11% higher than the saturated boiling heat transfer coefficients with an outlet vapour quality around 1.00, 13-15% higher than those measured with 5

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°C of outlet vapour super-heating, and 39-46% higher than those measured with 10 °C of outlet vapour super-heating. The slight decrease of the heat transfer coefficient with increasing vapour quality is probably due to dry-out inception in the upper part of the evaporator. The marked decrease of the heat transfer

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coefficient with vapour super-heating is due to the increase in the super-heating portion of the heat transfer surface, which is affected by gas single phase heat transfer coefficients one or two orders of magnitude lower than the two phase heat transfer coefficients in the boiling portion of the heat

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transfer surface.

The experimental saturated boiling heat transfer coefficients were compared against traditional

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equations for nucleate boiling, such as Cooper (1984) and Gorenflo (1993), and also against a new model specifically developed for refrigerant vaporisation inside BPHEs (Longo et al., 2015).

This new model consists of a non-dimensional equation based on the equivalent Reynolds number Reeq and the liquid Prandtl number PrL for the computation of the convective

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boiling heat transfer coefficient:

0.8

1/3

PrL

(10)

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hcb = 0.122 Φ (λL / dh) Reeq

(11)

PrL = µL cpL / λL

(12)

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Reeq = (G dh / µL) [(1 – Xm) + Xm(ρL / ρG)1/2]

where Φ is the enlargement factor of the corrugated plates (equal to the ratio between the actual area and the projected area of the plates).

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A new correlation, based on the Gorenflo (1993) equation, was developed for the computation of the nucleate boiling heat transfer coefficient:

hnb = 0.58 Φ h0 CRa F(p*) (q/q0)0.467

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(HFC32)

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h0 = 4000 W m-2 K-1

(13)

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CRa = (Ra / 0.4 µm)0.1333

F(p*) = 1.2 p*0.27 + [2.5 + 1 / (1 - p*)] p*

(14)

(15)

(16)

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h0 is the reference value of the heat transfer coefficient, p* the reduced pressure, and Ra the arithmetic mean roughness of the plates as defined in ISO4287/1. The final boiling heat transfer coefficient hboil is the maximum between the convective boiling heat

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transfer coefficient hcb, and the average nucleate boiling heat transfer coefficient hnb:

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hboil = MAX (hcb, hnb)

(17)

The heat transfer coefficients hboil, hcb, hnb are referred to the projected area of the plates. The absolute mean percentage deviation between calculated and experimental data is 19.1%, 32.1% and 4.7% for Cooper (1984) equation, Gorenflo (1993) equation, and Longo et al. (2015) model, respectively. Figure 6 shows the comparison between the experimental saturated boiling heat transfer coefficients and the calculated values by Longo et al. (2015).

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Figure 7 shows the saturated boiling frictional pressure drop against the kinetic energy per unit volume of the refrigerant flow computed by the homogeneous model:

KE/V = G2 / (2 ρm)

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(18)

The frictional pressure drop shows a linear dependence on the kinetic energy per unit volume of the refrigerant flow and therefore a quadratic dependence on the refrigerant mass flux. This linear

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relationship was already observed by Jassim et al. (2006) in adiabatic two-phase flow of HFC134a through a PHE. The following best fitting linear equation was derived from present experimental

∆pf [kPa] = 1.666 KE/V [J m-3]

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data:

(19)

around 5.3%.

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This correlation reproduces present experimental data with a mean absolute percentage deviation

The linear relationship between frictional pressure drop and the kinetic energy means a constant

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value for the friction factor

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f = ∆pf (dh / L) (2ρm / G2) (20)

vs. the refrigerant mass flux and therefore the Reynolds number as it occurs in the Moody diagram for single phase turbulent flow inside tube in the fully rough zone. Therefore, probably, for each plate corrugation and geometry it is possible to determine a specific constant to correlate frictional pressure drop and kinetic energy.

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HFC32 seems to be a very promising substitute for HFC410A in medium size chillers and heat pumps; therefore, it is interesting to compare its performance during vaporisation inside BPHEs to those of HFC410A. Figure 8 and 9 show the comparison between present HFC32 saturated boiling heat transfer coefficients and frictional pressure drops at 20 °C and those of HFC410A previously

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measured by Longo and Gasparella (2007) inside the same BPHE under the same operating conditions. HFC32 exhibits heat transfer coefficients 20-30% lower and frictional pressure drops 30-40% higher than those of HFC410A. This can be attributed mainly to the lower reduced pressure

3.

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of HFC32 with respect to HFC410A.

IR THERMOGRAPHY ANALYSIS

The measurements of heat transfer coefficients and pressure drops were complemented with the IR thermography analysis of the vaporisation process inside the BPHE. This analysis is aimed to

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investigate the heat transfer regimes during refrigerant vaporisation inside a BPHE in order to quantify the portion of heat transfer area affected by vapour super-heating (single-phase heat

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transfer) with respect to that working in boiling (two-phase heat transfer). In fact, in the real operating conditions of a BPHE evaporator inside a refrigerating machine, some degrees of vapour

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super-heating are requested at the outlet of the evaporator to prevent wet compression. The required vapour super-heating depends on the nature of the refrigerant (low, medium or high molecular weight), the type of compressor (alternative, scroll, screw, centrifugal), and the type of expansion device (thermostatic or electronic valve). In order to quantify the portion of the heat transfer surface affected by vapour super-heating the flank of the BPHE was filmed during the experimental tests by a IR thermo-camera (temperature uncertainty (k= 2) = ±0.1 °C in the temperature range 5 – 150 °C). Figures 10, 11, 12 and 13 show the results of the IR thermography carried out during HFC32 vaporisation tests at 10 and 20 °C of

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saturation temperature with a heat flux around 10 and 20 kWm-2 and different evaporator outlet conditions: the dotted line indicates the BPHE profile. During the saturated boiling tests (Xout = 0.80 and 1.00) the whole heat transfer surface works in two-phase heat transfer and it is near to saturation temperature (blue colour). At 5 °C of outlet vapour super-heating around 15-30% of the

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heat transfer surface (yellow-green area in the upper part of the BPHE) is affected by super-heating, whereas at 10 °C of outlet vapour super-heating this portion increases up to 40-50% (red area in the upper part of the BPHE). The results of the IR thermography analysis contribute to explain the great

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sensitivity of the heat transfer coefficients to the evaporator outlet conditions confirming that the dry-out phenomena produces a relevant degradation of the heat transfer performance of the BPHE

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evaporators. For this reasons the degree of vapour super-heating at the outlet of the evaporator must be limited at the minimum value compatible with a safe operation of the compressor and a stable operation of the expansion device. Present results suggest to not exceed 3-5 °C of vapour outlet

CONCLUSIONS

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super-heating in order to not compromise the evaporator efficiency.

This paper investigates the effects of heat flux, saturation temperature (pressure), and outlet

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conditions on HFC32 boiling heat transfer coefficient inside a BPHE. The heat transfer coefficients show weak sensitivity to saturation temperature (pressure) and great sensitivity to heat flux and outlet conditions. The saturated boiling heat transfer coefficients are 3946% higher than those with 10 °C of outlet vapour super-heating. The saturated boiling heat transfer coefficients are in fair agreement with a new model for refrigerant vaporisation inside BPHE (Longo et al., 2015). Furthermore, HFC32 exhibits heat transfer coefficients 20-30% lower and frictional pressure drop 30-40% higher than those of HFC410A. Finally, the IR thermography analysis confirms that the dryout phenomena produces a great degradation of the heat transfer performance of the BPHE evaporator and suggests to limit the

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degree of vapour super-heating at the outlet of the evaporator at the minimum value compatible

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with a safe operation of the compressor.

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ACKNOWLEDGMENTS

This research project was partially funded by: CariVerona Foundation, Verona, Italy, Ricerca Scientifica e Tecnologica 2013-2014: “Sviluppo di innovativi processi a ridotto impatto ambientale per la conservazione e distribuzione a

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bassa temperatura delle derrate alimentari a salvaguardia della salute”.

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REFERENCES

ANSI / ASHRAE 34, 2013. Designation and Safety Classification of Refrigerants. Bella, B., Kaemmer, N., Brignoli, R., Zilio, C., 2014. Energy Efficiency of a Chiller Using R410A or R32. Proc. 15th International Refrigeration and Air Conditioning Conference at

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Purdue, West Lafayette, IN, USA. Cooper, M.G., 1984. Heat flows rates in saturated pool boiling - a wide ranging examination using reduced properties. In: Advanced in Heat Transfer. Academic Press, Orlando, Florida, pp. 157-239. Dutto, T., Blaise, J.C., Benedic, T., 1991. Performances of brazed plate heat exchanger set in heat

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pump. Proc. 18th Int. Congr. Refrigeration, Montreal, Canada, 1284-1288.

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Fernando, P., Palm, B., Lundqvist, P., Granryd, E., 2004. Propane heat pump with low refrigerant charge: design and laboratory tests. Int. J. Refrigeration 27, 761–773. Gorenflo, D., 1993. Pool boiling. VDI Heat Atlas, Dusseldorf, Germany, Ha1-25. Kline, S.J., McClintock, F.A., 1953. Describing uncertainties in single-sample experiments. Mech.Eng. 75, 3–8.

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Jassim, E.W., Newell, T.A., Chato, J.C., 2006. Refrigerant pressure drop in chevron and bumpy style flat plate heat exchangers, Exp. Therm. and Fluid Science, 30, 213-222.

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Jung, Y.K., Kim, H.J., Lee, H.S., Lee, Y.S., 2014. Heat Transfer and Pressure Drop Characteristics of Plate Heat Exchanger of R22 and R32. Proc. Twenty-fourth International Ocean and Polar

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Engineering Conference, Busan, Korea. Lee, H.S., Kim, H.J., Cha, S.W., Moon, D.S., 2014. Performance Characteristics of Seawater Heat Pump Using R32/R152a Mixed Refrigerant. Proc. Twenty-fourth International Ocean and Polar Engineering Conference, Busan, Korea. Longo, G.A., Gasparella, A., 2007. Heat transfer and pressure drop during HFC refrigerant vaporisation inside a brazed plate heat exchanger. Int. J. Heat Mass Transfer 50, 5194–5203. Longo, G.A., Gasparella, A., 2007. R410A vaporisation inside a commercial brazed plate heat exchanger. Experimental Thermal and Fluid Science 32, 107–116.

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Longo, G.A, Mancin, S., Righetti, G., Zilio, C., 2015. A new model for refrigerant boiling inside a Brazed Plate Heat Exchanger (BPHE). Proc. 24th International Congress of Refrigeration, Yokohama, Japan. Mancin, S., Del Col, D., Rossetto, L., 2013. R32 partial condensation inside a brazed plate heat

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exchanger. Int. J. Refrigeration 36, 601–611. Palm, B., 2007. Refrigeration systems with minimum charge of refrigerant. Appl. Therm. Eng. 27, 1693-1701.

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Palmer, S.C., Vance Payne, W., Domanski, P.A., 2000. Evaporation and Condensation Heat Transfer Performance of Flammable Refrigerants in a Brazed Plate Heat Exchanger. NIST. IR

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6541.

Regulation (EC) No 842/2006 of the European Parliament and of the Council of 17 May 2006 on certain fluorinated greenhouse gases.

Regulation (EU) No 517/2014 of the European Parliament and of the Council of 16 April 2014 on fluorinated greenhouse gases and repealing Regulation (EC) No 842/2006.

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United Nations. Montreal Protocol on Substances That Deplete the Ozone Layer. United Nations

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Environment Program, Ozone Secretariat, New York, NY, USA, 1987.

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Table 1. Specification of the different measuring devices Uncertainty (k=2)

Range

T-type thermocouples

0.1 K

-20/80 °C

T-type thermopiles

0.05 K

-20/80 °C

Abs. pressure transducers

0.075% f.s.

Diff. pressure transducers

0.075% f.s.

Coriolis effect flow meters

0.1%

Magnetic flow meters

0.15% f.s.

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0/0.3 MPa

0/300 kg h-1 100/1200 l h-1

± 2.7 µV

0 / 100 mV

± 0.1 °C

-20 / 250 °C

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Thermo camera

0/3.0 MPa

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Data logger

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Devices

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Table 2. Geometrical characteristics of the BPHE evaporator.

Parameter

Measure / Type 278.0

Plate width W(mm)

72.0

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Fluid flow plate length L (mm)

Area of the plate A(m2)

0.020

Chevron

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Corrugation type Angle of the corrugation β(°)

Corrugation pitch P(mm) Plate roughness Ra(µm)

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Corrugation dept b(mm)

65

2.0

8.0 0.4 1.0

Total number of plates

10

Number of effective plates

8

Channels on refrigerant side

4

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5

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Table 3. Operating conditions during experimental tests. Tsat(°C)

psat(MPa)

Xin

Xout

∆Tsup(°C)

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5.0–20.1

0.95–2.30

0.19–0.33

0.79–0.98

4.7–10.9

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Gr(kg m-2s-1)

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Runs

9.5–29.1

Gw(kg m-2s-1)

q(kW m-2)

55.6–190.0

5.2–24.0

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Captions to the figures

Figure 1. Schematic view of the experimental test rig. Figure 2. Schematic view of the plate.

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Figure 3. Average heat transfer coefficient on refrigerant side vs. heat flux at 5 °C of saturation temperature.

Figure 4. Average heat transfer coefficient on refrigerant side vs. heat flux at 10 °C of saturation

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Figure 5. Average heat transfer coefficient on refrigerant side vs. heat flux at 20 °C of saturation

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Figure 6. Comparison between experimental and calculated heat transfer coefficient. Figure 7. Saturated boiling frictional pressure drop vs. kinetic energy per unit volume. Figure 8. Comparison between HFC32 and HFC410A saturated boiling heat transfer coefficients inside a BPHE at 20 °C of saturation temperature.

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Figure 9. Comparison between HFC32 and HFC410A saturated boiling frictional pressure drop inside a BPHE at 20 °C of saturation temperature.

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Figure 10. IR thermography during vaporisation tests at 10 °C of saturation temperature with 10 kWm-2 of heat flux.

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Figure 11. IR thermography during vaporisation tests at 10 °C of saturation temperature with 20 kWm-2 of heat flux.

Figure 12. IR thermography during vaporisation tests at 20 °C of saturation temperature with 10 kWm-2 of heat flux.

Figure 13. IR thermography during vaporisation tests at 20 °C of saturation temperature with 20 kWm-2 of heat flux.

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ACCEPTED MANUSCRIPT HFC32 VAPORISATION INSIDE A BRAZED PLATE HEAT EXCHANGER (BPHE): EXPERIMENTAL MEASUREMENTS AND IR THERMOGRAPHY ANALYSIS

HIGHLIGHTS

This paper investigates HFC32 vaporisation inside a BPHE

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HFC32 exhibits heat transfer coefficients slightly lower than those of HFC410A

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HFC32 exhibits frictional pressure drop slightly higher than those of HFC410A

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The experimental measurements are complemented with a IR thermography analysis

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