Thermal analysis of oil sump and compression unit in a rotary compressor

Applied Thermal Engineering 164 (2020) 114465

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Thermal analysis of oil sump and compression unit in a rotary compressor Hongyan Shi, Jianhua Wu

⁎

T

School of Energy and Power Engineering, Xi′an Jiaotong University, Xi′an 710049, PR China

H I GH L IG H T S

sump and compression unit were thermally simulated by 3D-coupled CFD model. • Oil of oil sump and compression unit vary at different directions. • Temperatures sump dissipates more heat to compression unit than that to compressor shell. • Oil unit absorbs more heat from discharge gas than that from oil sump. • Compression • An adiabatic coating on the compression unit could reduce suction superheat degree.

A R T I C LE I N FO

A B S T R A C T

Keywords: Rotary compressor Oil sump Compression unit CFD fluid-solid coupled model Temperature distribution Heat transfer mechanism

Thermal analysis of the oil sump and compression unit in a hermetic rotary compressor is important for enhancing its performance and reliability. In this study, the temperature distribution, heat-transfer mechanism, and heat balance of the oil sump and compression unit in a rotary compressor were analyzed using the 3D computational fluid dynamics (CFD) fluid-solid coupled model. Considering the practical characteristics of the rotary compressor, some special treatments are proposed for the simulation model. Instead of using empirical formulations, heat-transfer coefficients in the upside of the main bearing and compressor shell were obtained using the 3D CFD simulation model; the dynamic process in the suction chamber and compression chamber was considered to be static and the oil film in the inner wall of the cylinder, rib effect of the A-frame structure at the bottom of the compressor shell, and friction loss in the side of the sliding vane were considered. The simulation results showed good agreement with experimental results. We expect that these analytical results would be of great use in further understanding the temperature distribution and heat-transfer mechanism of oil sumps and compression units and also provide guidelines for improving the performance and reliability of rotary compressors.

1. Introduction At present, hermetic rotary compressors with high-pressure shells are being used universally in room air conditioners for their high reliability, efficiency, and compact structure. The heat-transfer characteristics and temperature distribution of the oil sump and compression unit (which mainly includes the cylinder, roller, main bearing, sub bearing, muffler, sliding vane, and crankshaft) in a rotary compressor have a significant influence on its performance and reliability; they affect several parameters, such as the superheat of the suction gas for cooling capacity, thermal deformation, oil properties, friction loss, and leakage loss. Therefore, it is essential to achieve a good understanding of the heat-transfer mechanism of and temperature distribution in the oil sump and compression unit of a rotary compressor.

⁎

Several reports are available on the thermal analysis of rotary compressors. Yanagisawa et al. [1] divided a cylinder into many elements to numerically study the heat of the gas in the suction chamber and the heat-transfer coefficient was set at the interface of the oil sump and compression unit. Padhy et al. [2] simulated the heat-transfer characteristics of rotary compressors using the lumped parameter method; all parts connected to the gas or oil were set as known heattransfer coefficient boundary conditions and friction loss was loaded on the friction surface. Liu [3] modelled the temperature distribution of high-speed rotary compressors using the lumped parameter method and proposed a new heat-transfer coefficient model for the cylinder inner wall; however, in this study, it was assumed that the oil-sump temperature and cylinder temperature were uniform. Ishii [4] numerically simulated heat transfer in a cylinder and suggested that most of the heat

Corresponding author at: Postal address: No. 28, Xian-ning West Rd., Xi′an City 710049, Shaanxi, PR China. E-mail address: [email protected] (J. Wu).

https://doi.org/10.1016/j.applthermaleng.2019.114465 Received 10 March 2019; Received in revised form 24 September 2019; Accepted 29 September 2019 Available online 30 September 2019 1359-4311/ © 2019 Published by Elsevier Ltd.

Applied Thermal Engineering 164 (2020) 114465

H. Shi and J. Wu

Nomenclature T T P RPM n θ φ ϕ h h′ H R D V A δ ṁ Q Cp ρ β

g Nu d Re Pr λ ω μ u

temperature (℃) temperature (K) pressure (MPa) rotation speed (r/min) rotation speed (r/min) crank angle (rad) crank angle (rad) crank angle (°) heat transfer coefficient (W·m−2·K−1) equivalent heat transfer coefficient (W·m−2·K−1) the height of cylinder (mm) radius (mm) diameter (mm) stroke volume (m3) surface area (m2) thickness of oil film (um) mass flow rate (kg·h−1) heat (W) average specific heat capacity (J kg−1∙ K−1) density (kg∙m−3) thermal expansion coefficient

gravity acceleration (9.8 m·s−2) nuselt number suction tube diameter (mm) Renolds number Prandtl number thermal conductivity (W·m−1·K−1) rotating speed of crankshaft (rad·s−1) the friction coefficient speed (m·s−1)

Subscript e s c a ave gc gs hy cy o vc

evaporating suction compression ambient average value compression chamber suction chamber hydraulic cylinder oil sliding vane

of the A-frame structure at the bottom of the compressor shell, and friction loss in the side of the sliding vane were considered. Meanwhile, temperatures at different places in the cylinder and oil sump were measured to verify the accuracy of the simulation results. Further, a methodology is proposed to prevent gas overheating in the suction chamber. The results presented in this report would be useful for further understanding the temperature distribution in and heat-transfer mechanism of oil sumps and compression unit s and also provide guidelines for improving the performance and reliability of rotary compressors.

was transmitted from thrust bearings; Colburn′s empirical analogy was used to calculate the heat-transfer coefficient. However, these studies focused only on some parts of the compressor and compression unit, there are no reports on the combined effect of the oil sump and compression unit. In the past, most studies on temperature distribution in a compressor employed the lumped parameter method. Liu et al. [5] divided the compressor into eight parts while Chen et al. [6] divided an overall scroll compressor into 9 parts and simulated them using the lumped parameter method. Ooi [7] divided a reciprocating compressor into 46 parts to study it. In recent years, with the development of computer technology, methods with improved calculation accuracy, such as computational fluid dynamics (CFD) and finite element method (FEM), are being employed to study compressors. The CFD method has been widely used to study the temperature distribution and fluid distribution in reciprocating compressors [8–16]. In addition, Hsieh [17] studied temperature distribution in a screw compressor using a combination of self-programming and CFD methods. Shen [18] modelled temperature distribution in compressed natural gas (CNG) compressors by FEM and friction loss was loaded on the friction surface. Ooi et al. [19] also used the CFD method to study fluid flow and heat transfer in the working chamber of a scroll compressor. Evandro [20] used a 2D unsteady CFD model to simulate gas-temperature distribution in the working chamber of a scroll compressor. Compared to R410A, R32 is a better substitution refrigerant for its zero Ozone Depletion Potential (ODP) and lower Global Warming Potential (GWP), and advantageous thermal properties. However, the rather higher discharge temperature of R32 than R22 and R410A would result in more performance and reliability problems (the oil viscosity could not satisfy the lubricant requirement and then increase friction loss; the great thermal deformation occurred in the cylinder, and the motor being burned out, etc.). In this study, the 3D CFD fluid-solid coupled method was employed to evaluate the temperature distribution and heat-transfer mechanism of the oil sump and compression unit in a R32 rotary compressor. In the simulation model, instead of using empirical formulations, heat-transfer coefficients in the upside of the main bearing and compressor shell were obtained using the 3D CFD simulation model; the dynamic process in the suction chamber and compression chamber was considered to be static and finally, the oil film in the inner wall of the cylinder, rib effect

2. Model 2.1. Heat-transfer mechanism of an oil sump and compression unit In a rotary compressor, oil flows to the lubricant parts through an oil hole at the bottom of the crankshaft [21] (Fig. 1, red arrows illustrate

Fig. 1. Schematic view of a rotary compressor. 2

Applied Thermal Engineering 164 (2020) 114465

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Fig. 3(a) and (b), respectively. The first inlet of the oil sump is located at the top (Fig. 3(a), inlet I) and the bottom of the sub bearing, which connects with the outer surface of the crankshaft, is the outlet of the spiral groove (Fig. 3(a), inlet II). The bottom of the centre hole in the crankshaft is the outlet of the oil sump. The corresponding mesh model of the oil sump and compression unit is shown in Fig. 3(c). It comprises 54.3 million unstructured and approximately tetrahedral grids. Meshes in places where the temperature or velocity changed greatly or where the area was relatively small, such as the oil sump inlet and outlet, were refined. Meanwhile, grids in places with small gradients should be meshed sparsely for higher computational efficiency. The grid independence verification is shown in Table 2 (the test points (1)–(8) in Fig. 9 were chosen as the comparisons).

the oil path). Some part of the oil flows into the main bearing spiral groove and then flows out from the top of the spiral groove back to the oil sump; heat generated due to friction in the main bearing would be taken to the oil sump by this oil. In addition, some part of the oil flows into the spiral groove in the sub bearing and flows back to the oil sump through the outlet of the spiral groove; in this case as well, frictional heat generated in the sub bearing is transported to the oil sump by the oil flowing back to it. The rest of the oil flows into a cavity (formed by the roller, crankshaft, and bearing) to lubricate the eccentric bearing and absorb its frictional heat; this oil then leaks into the working chamber through the end face clearance of the roller and is discharged from the muffler after mixing with the discharge gas. Finally, the oil is thrown to the inner wall of the compressor shell due to centrifugal action of the rotor and flows to the oil sump along the compressor shell. In the sump, oil flows from the top to the bottom, during which process heat is dissipated to the compressor shell and compression unit. A compressor compression unit is mainly composed of a cylinder, roller, main bearing, sub bearing, muffler, sliding vane, and crankshaft. The outside of the compression unit is covered by the oil sump and discharge gas. The innards of the compression unit, i.e., the working chamber, are divided into a suction chamber and compression chamber by a roller and sliding vane, which are in contact with the refrigerant. The temperature of the refrigerant in the suction chamber and suction tube is lower than the oil-sump temperature and hence refrigerant in the suction chamber and suction tube would be heated by the oil sump through the cylinder, main bearing, and sub bearing, as well as by the oil leaking through the roller. In addition, high-temperature discharge gas present in the muffler would heat the refrigerant through the main bearing. The main heat-flow diagram of the rotary compressor is shown in Fig. 2.

2.3. Boundary conditions A coupled boundary condition was specified for all the surfaces in contact with the oil sump, such as the cylinder, bearings, and crankshaft. Some data needed in the compressor shell-air zone model (in Section 2.3.3) are obtained from the compressor shell temperature test points 1–7 (shown in Fig. 9). Besides, the inlet gas temperature of the muffler CFD model (in Section 2.3.5) is also obtained from the tested point (8) (shown in Fig. 9); other boundary-treatment methods are as follows. Simulation conditions of GX, ARI, and ASHRAE/T1 (shown in Table 3) were adopted; here, the GX condition represents a high-efficiency condition designed to match the air conditioner with a compressor performance (COP) of 3.4–3.6 [22]. 2.3.1. Interior of the compression unit In the simulation model, the surface of the working chamber (cylinder inner wall, downside of the main bearing, and upside of the sub bearing) is divided into 4 parts in the circumferential direction. In the simulation model, the average heat-transfer coefficient and gas temperature are set in every part. The calculation procedure is described below. 1) Calculation of average gas temperature in the working chamber. According to the values of te, ts, and tc and considering the influence of the discharge valve, backflow of gas in the clearance volume, and

2.2. The physical model and mesh model The heat-transfer mechanism of and temperature distribution in the oil sump and compression unit in a hermetic rotary compressor (shown in Fig. 1) with a high-pressure shell, single cylinder, and single discharge port (in the main bearing) will be analysed in this study. The main parameters of the compressor are listed in Table 1. As the working process in the working chamber is complicated, to simulate the oil sump and compression unit efficiently, some simplifications were made in the model; they are as follows: 1) The muffler was not modelled in the oil sump and compression unit coupled model. Instead, the muffler was modelled and simulated independently by CFD and heat-transfer coefficient between the discharge gas in the muffler and main bearing was then calculated. 2) As shaft temperature corresponding to the vertical part of the bearings (main and sub) is approximately the same as the temperature of the bearings, the shaft was entirely constructed with the vertical part of the bearings. 3) To reduce simulation time, the dynamic process in the suction chamber and compression chamber was assumed to be static and the known average heat-transfer coefficient and average gas temperature were used for the internal surface of the working chamber. 4) In the physical model, the A-frame at the bottom of the compressor, which has a better thermal effect on heat dissipation in the compressor shell, was neglected and an equivalent heat-transfer coefficient was used in the place of this compressor shell. 5) The oil level differs under different operating conditions. In this study, the oil level was consumed in the upside of main bearing (commonly used level). 6) As the roller rotates and rolls with the crankshaft, and its temperature at different places is almost uniform. In this model, it was consumed as an adiabatic part, and was not modelled.

Fig. 2. Heat-flow diagram.

Physical models of the oil sump and compression unit are shown in 3

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Table 1 Basic parameters of the studied rotary compressor. Dcy (mm)

Hcy (mm)

Vth (cm3)

Dshell (mm)

Suction tube diameter (mm)

Discharge tube diameter (mm)

Shell internal pressure

50

19

19

118

16.4

8.5

Pc

leakage loss (from the end face of roller, end face of sliding vane, side face of sliding vane, and radial clearance between roller and cylinder), temperature variation of the gas in the compression chamber and suction chamber was evaluated under different conditions (GX, ARI, ASHRAE\T1). The curve generated in the GX condition is presented in Fig. 4. The gas average temperature in circumference direction is calculated according to the above data in this section and the formula proposed by Yanagisawa et.al [1]:

Tave =

1 ( 2π

∫0

φ

Tgc (θ) dθ +

∫φ

2π

Tgs (θ) dθ)

Table 2 Grid independence verification.

have

∫0

φ

hgc (θ) dθ +

∫φ

2π

hgs (θ) dθ)

Point (2) (℃)

Point (3) (℃)

Point (4) (℃)

Point (5) (℃)

Point (6) (℃)

2,056,421 5,430,122 8,521,471

81.5 80.2 80.9

75.1 71.5 71.8

72.7 73.9 73.2

62.1 65.8 65.1

73.7 71.0 71.6

76.8 74.0 74.1

h′ = 1/(

δ 1 + ) λo have

(4)

where δ is the thickness of the oil film, λo is its thermal conductivity, and have is the average convective heat-transfer coefficient between the cylinder and refrigerant (shown in Eq. (3)).

(2)

where, к is the specific heat ratio and u = 2ωRc, Dhy = 4 V(α)/A(α), Re = uDhy/υ, and Pr = Cpμ/λ; RCr1 is the radiation of the cylinder, rave Rave is average radiation of the cylinder and piston, DhDh is hydraulic diameter, V(α) is stroke volume, A(α)A(α ) is the inner surface area of chamber, and λ, u, υ, μ, and Cp are properties of the fluid in the inner chamber of the cylinder. The suction chamber volume, compression chamber volume, and Dh vary with respect to the rotation of the crankshaft. Furthermore, the density, thermal conductivity, and viscosity of the refrigerant in the cylinder vary with respect to variations in temperature and pressure. In one cycle, the average heat-transfer coefficient can be represented as follows:

1 = ( 2π

Point (1) (℃)

where: hgc (θ) is the heat-transfer coefficient in the compression chamber, hgs (θ) is the heat-transfer coefficient in the suction chamber, θ is the rotation angle of the crankshaft, and φ is the angle at a certain point in the cylinder. The results in different conditions, calculated using Eq. (3), are plotted in Fig. 6. 3) The impact of oil film on the inner wall of cylinder. In practice, during the operation of a rotary compressor, oil leaks into working chamber from the interior of the roller through the clearance of roller end face. During the rolling of the roller, oil film with tens of micrometer thick in thickness (approximately equals to the minimum radial clearance between the roller and cylinder) evenly attaches to the cylinder inner wall. Thermal conductivity of the oil is much lesser than that of the cylinder and hence its impact on heat transfer between the refrigerant and cylinder cannot be neglected. The equivalent heat-transfer coefficient (h’) between the cylinder and refrigerant can be represented as follows:

(1)

where Tgs (θ) is gas temperature in the suction chamber, Tgc (θ) is gas temperature in the compression chamber, θ is the crank angle, and φ is the angle at a certain point on the cylinder inner wall. The obtained results are depicted in Fig. 5. 2) The calculation of average heat transfer coefficient in the surface of working chamber. The surface of working chamber in the same place is set as the same boundary condition. Now, the boundary condition can be obtained by the calculation of cylinder. The complex motion of the refrigerant in the working chamber results in a heat exchange more complex than that in the suction tube. After summarising studies conducted in the past [1,23–25], Liu [3] used a new heat-exchange equation (shown in Eq. (2)) to calculate the heattransfer coefficient between the inner wall of the cylinder and gas in the cylinder, which is also used in this study:

Dhy κ hc = 0.025 Re 0.8Pr 0.4 (1.0 + 1.77 ) Rave Dh

Number of elements

2.3.2. Oil sump Inlet I (in Fig. 3(a)) is set as mass-flow-inlet condition and the mass flow rate ṁ inletI represents oil discharged along with the refrigerant from the muffler and oil coming out from the spiral groove of the main bearing. Considering that the oil participates in heat exchange with the compressor shell before dropping to inlet I, inlet oil temperature, TinletI, is lower than that of the discharge gas; TinletI is calculated as follows,

(3)

hshell ·Ashell ·(Ta − Tshell ) = ṁ d ·Cpd − d′ ·(Td − Td′)

(5)

ṁ d ·Cpd′− inletI ·(Td′ − TinletI ) = ṁ mb ·Cpmb − inletI ·(TinletI − Tmb (out ))

(6)

Qf mb = ṁ mb ·Cpmb (in) − mb (out ) ·(Tmb(in) − Tmb(out ))

(7)

Fig. 3. Physical and mesh model of the oil sump and compression unit in a rotary compressor, (a) physical model of oil sump, (b) physical model of compression unit, (c) grid model of oil sump- compression unit. 4

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Table 3 Simulation condition. Condition

tc (℃)

Pc (MPa)

te (℃)

Pe (MPa)

ts (℃)

tval (℃)

ta (℃)

RPM (r/min)

Motor efficiency (%)

GX ARI ASHRAE/T1

46.0 54.4 54.4

2.862 3.473 3.473

10.0 7.2 7.2

1.107 1.018 1.018

18.0 18.3 35.0

41.0 46.1 46.1

35.0 35.0 35.0

2880 2880 2880

84 84 84

120

Gas in discharge chamber

ṁ inletII = ṁ − ṁ inletI

105

Gas in suction chamber

where hshell is the heat-transfer coefficient between the compressor shell and ambient environment, which can be calculated by the CFD model of the compressor shell in an air domain (described in Section 2.3.3). Ta is the ambient temperature, Tshell is the compressor-shell temperature, Td is the discharge temperature, Td’ is oil temperature in inlet I before mixing with oil from the spiral groove in the main bearing, and TinletI and TinletII represent oil temperatures in inlet I and II, respectively. Tmb(in)’ and Tmb(out) are oil temperatures in the inlet and outlet of the spiral groove in the main bearing, respectively. Similarly, Tsb(in)’ and Tsb(out) represent oil temperatures in the inlet and outlet of the spiral groove in the sub bearing, respectively. ṁ d , ṁ mb , and ṁ indicate oil mass in the outlet of the muffler, oil mass in the spiral groove of the main bearing, and total oil mass flowing to the oil sump, respectively; these values were sourced from [21]. Cpd-d’, Cpd’-inletI, CpinletI-mb, Cpmb(in)mb(out), and Cpsb(in)-sb(out) represent average specific volumes of the oil at temperature differences between Td and Td’, Td’ and TinletI, TinletI and Tmb(out), Tmb(in) and Tmb(out), and Tsb(in)and Tsb(out) respectively; these values were obtained from [26]. Qf mb and Qf sb are frictional forces in the main bearing and sub bearing, respectively, which can be calculated using the following friction-loss equations [27],

t( )

90 75 60 45 30 15

0

60

120 180 240 300 360

φ

( )

Fig. 4. Gas temperature in the suction and compression chambers under GX conditions.

Qf

mb

Qf

sb

= Mmb ωs

= Msb ωs

(10)

(11) (12)

where Mmb and Mmb are the torques of the main bearing and sub bearing, respectively, and ωs is the rotational speed. The shape and size of the oil sump are directly affected by the compression unit and compressor shell. Oil velocity in the sump is relatively low except at the bottom (which is adjacent to the crankshaft) and hence a laminar model was chosen to model oil flow in the oil sump and the Boussinesq model (Eq. (13)) was selected to represent fluid density.

ASHARE/T1 condition

(ρ − ρ0 ) g ≈ −ρ0 β (T − T0) g

1500

GX condition

-1

1350

ARI condition

1200

here, ρ0 is the set density of oil and T0 is the operating temperature. In addition, the under surface of the crankshaft was set as a rotational wall and rotational speed was equal to the rotational speed of the compressor.

have (W·m ·K )

1650

-2

Fig. 5. Average gas temperature (tave) along the circumference under different conditions.

1050 900

2.3.3. Compressor shell In previous simulation models, the heat-transfer coefficient of the compressor shell was often obtained by empirical formulations. To further increase the accuracy of the obtained results, in later studies, the heat-transfer coefficient was evaluated using a coupled CFD model for the compressor shell-air zone [28]. Due to the action of gravity and different shell temperatures (corresponding to the oil sump, compression unit, and motor), the heat-transfer coefficient of the compressor shell varies with respect to height. To distinguish these differences, the compressor shell was divided into 7 parts and their heat-transfer coefficients were calculated. The A-frame at the bottom of the compressor (shown in Fig. 7(a)) can improve heat transfer for the reason that the heat transfer area is as 4 times large as the area shown in Fig. 7(b). To simplify the simulation model, the A-frame was not modelled but the heat-transfer coefficient of the bottom of the compressor (red part in Fig. 7(b)) is large (calculated according to the total transfer heat, area ratio, and rib efficiency

750 600 0

60

120 180 240 300 360 φ( )

Fig. 6. Average heat-transfer coefficient (have) at different points on the cylinder inner wall.

ṁ inletI = ṁ mb + ṁ d

(8)

Inlet II (in Fig. 3(a)) also follows a mass-flow-inlet condition and its mass flow rate ṁ inletII and temperature tsb(out) are determined using the following equations,

Qf sb = ṁ inletII ·Cpsb(in) − sb(out ) ·(Tsb(in) − Tsb(out ))

(13)

(9) 5

Applied Thermal Engineering 164 (2020) 114465

H. Shi and J. Wu

2.3.7. Fundamental equations Due to the steady-state operation of the compressor, and zero heat resources in the uniform solid parts of the simulation model, so the heat conduction in the solid parts in a 3D Cartesian coordinate system was simplified as Eq. (17):

∂ ∂T ∂ ∂T ∂ ∂T (λ x )+ (λ y )+ (λ z )=0 ∂x ∂x ∂y ∂y ∂z ∂z

(17)

where, T is temperature, λx, λy and λz represent thermal conductivity of the solid in the x, y, and z directions, respectively, and λx = λy = λz = constant. Under steady state condition, oil flow in the oil sump is laminar with a viscous force. Therefore, the viscous-laminar model in ANSYS Fluent was chosen to analyse the oil sump. As the liquid fluid density changes a little with temperature, the fluid density ρ and kinematic viscosity v are consumed constant. The governing equations of this model are presented as follows: Mass conservation equation:

Fig. 7. Treatment of A-frame at the bottom of oil sump.

[2]).

⇀

λ h = 0.023 Re0.8Pre 0.4 d

Momentum conservation equation: ⇀ ⇀ ⇀ 1 (u ·∇) u = − ∇p + v∇2 u + f ρ ⇀

(14)

⇀ ⇀ ⇀ ⇀ ⇀ 1⇀ ⇀ ∇ ⎡ρu (e + u ·u )⎤ = ∇ ·(Σ·u ) + u ·ρf 2 ⎣ ⎦

(20)

⇀

where, u is velocity vector, ρ is density, v is kinematic viscosity, p is ⇀

pressure; f is mass force vector of per unit mass, of which fx = fy = 0, fz = −9.8 (m2∙s−1) (the coordinate system is the same with Fig. 9); e is internal energy of per unit mass.

2.3.5. Main bearing and muffler The upper surface of the compression unit (upside of main bearing) is divided into three parts. Zone-1 is the horizontal part covered by the muffler, where the heat-transfer coefficient would be obtained by the muffler-refrigerant CFD model. Zone-2 is the horizontal part not covered by the muffler; as this gas velocity is lower than gas velocity in the muffler, heat-transfer coefficient on its outer surface is about 20% of the coefficient in zone-1 (calculated according to the volume ratio). Zone-3 is the vertical part of the main bearing; the heat-transfer coefficient on its outer surface was calculated according to the equation postulated for a swept plate (Eq. (15)).

3. Experiment apparatus To achieve a clear understanding of temperature distribution in the oil sump and compression unit, we conducted temperature tests on the compressor performance test platform. From the experiment, the needed boundary conditions of the simulation model and data to validate it were obtained. 3.1. The compressor used in the experiment

(15)

The rotary compressor with high-pressure shell, single cylinder, and single discharge port (in the main bearing) used in this experiment is shown in Fig. 8 (the main parameters are shown in Table 1). Free convective heat transfer occurred on the outside of the compressor shell and R32 and POE oil were used in the test system.

To reduce computation time, the muffler and the discharge gas in it were modelled independently by CFD; this model was composed of a refrigerant zone, upper surface of the main bearing, the muffler, inlet of the muffler, and outlet of the muffler. Heat-transfer coefficient between the discharge gas in the muffler and main bearing (hmuf-mb) was calculated. Therefore, the muffler was no longer modelled in the coupled oil sump- compression unit model; instead, the heat-transfer coefficient and gas temperature were set at the upper surface of the main bearing. The simulated results under different conditions are shown in Table 4.

3.2. Temperature sensors and test points Temperature was measured at certain points in the oil sump and inner wall of the cylinder along with the suction gas temperature in the suction tube, discharge temperature in the discharge muffler, and outside temperature of the compressor shell. As described in other studies, thermocouples were used to measure these temperatures [1–3]. As fluid velocities in the suction port, bottom of the oil sump, and outlet of the muffler were higher than those at other parts, to prevent the thermocouples from being washed away by the fluid, sheathed thermocouples

2.3.6. Sliding vane Friction loss-1 comes from the top of the sliding vane and it is mainly transferred to the roller, while loss-2 from the side of the sliding vane is transferred to the cylinder and oil sump. In this simulation model, loss-2 was considered and it was loaded on the side of sliding vane as a heat resource. Side friction loss Lsi was calculated as follows [27]:

Lsi = μ vc |u vc|(|Fn1| + |Fn2|)

(19)

Energy conservation equation:

where λ is the thermal conductivity of the refrigerant and d is the suction tube diameter.

Nu = 0.664Re1/2 Pr 1/3

(18)

∇·u = 0

2.3.4. Suction tube The refrigerant is drawn into the suction chamber through the suction tube. Convective heat transfer occurs between the refrigerant and the inner wall of the suction tube; in this case, the Dittus-Boelter [29] equation is adopted to calculate the heat-transfer coefficient h,

Table 4 The simulated results of muffler under different conditions.

(16)

Condition

where, μvc is the friction coefficient of the side sliding vane, uvc is the reciprocating speed of the sliding vane, and Fn1 and Fn2 are the reaction forces of the sliding vane stressed by the sliding vane slot.

hmuf-mb (W∙m

6

−2

−1

∙K

)

GX

ARI

ASHRAE/T1

559

494

459

Applied Thermal Engineering 164 (2020) 114465

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The thermocouples were calibrated reciprocally and the differences among them were less than 1 ℃. The compressor was operated three times under each condition and in each cycle, the compressor operated steadily for 30 min. The uncertainties under the three conditions are all less than 0.5℃ (calculated with Eq. (21)). The test data presented in this paper are the average values of the three cycles.

Δ = max(x − x i )

(21)

where, Δ is the uncertainty, x is the average of the tested values, xi is the tested value in every time. 4. Results and discussion From the comparison in Table 5, it can be inferred that the simulation model could accurately represent the compressor; deviations between simulated and experimental data were less than 3 ℃. The simulated oil-sump temperatures were larger than the experimental values, while the simulated cylinder temperatures were lower than the experimental values. This may be because better heat transfer occurs at the interface of the oil sump and compression unit due to compressor vibration; however, these vibrations were not considered in the simulation model. Besides, the average heat transfer coefficient and average gas temperature set on the working chamber surface, and the simplification of the simulation model all bring deviations between the simulation and the experiment.

Fig. 8. The rotary compressor used in the experiment.

with a measuring error of ± 0.5 ℃ were required in these places. In other places, T-type thermocouples with a measuring error of ± 0.3 ℃ were used. In order to test the points temperature of (4)–(6) shown in Fig. 9, three holes (the diameters are about 2 mm) were drilled in the cylinder, and the T-type thermocouples were pasted in the tested positions by superglue, and then the holes were also filled with the superglue. The sheathed thermocouples in oil sump were directly fixed on the compressor shell by welding. The sheathed thermocouples in suction tube and discharge port were fixed in the test points by welding. At last these thermocouple wires were drawn out from the holes drilled on the compressor shell, and these holes were also sealed with superglue.

4.1. Analysis of temperature distribution of the oil sump It can be inferred from Table 4 that temperature distribution in the oil sump was uniform under different conditions. As an example, temperature-distribution contours of the oil sump under GX conditions are presented in Figs. 10–12. From Fig. 10(a) and (b), it can be deduced that oil temperature was lower near the suction tube than in other places and in a place opposite the suction tube, temperature decreased from the top of the oil sump to the bottom. In addition, as oil coming from the spiral groove of the sub bearing absorbs frictional heat produced by the sub bearing, near the outlet of oil sump, oil temperature is a little higher than the temperature at other places at the bottom of the sump. Fig. 11(a) and (b) show the longitudinal section views of the oil sump cut from the middle of the sliding vane and at a position 90° with

3.3. Test procedure and conditions The temperatures of the oil sump and compression unit were measured under GX, ARI, and ASHRAE/T1 conditions (shown in Table 2).

Fig. 9. The arrangement of temperature test points of compressor, (1)-upside of the oil sump, (2)-middle of the oil sump, (3)-bottom of the oil sump, (4)-inner wall of the cylinder close to the suction port, (5)-inner wall of the cylinder opposite the suction port, (6)-inner wall of the cylinder close to the discharge port, (7)-gas in the suction tube, and (8)-gas in the discharge port. 7

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Table 5 The comparison between experimental results (exp) and simulated results (sim). Conditions

GX (℃)

ARI (℃)

ASHRAE/T1 (℃)

Test points

Exp

Sim

Sim-exp

Exp

Sim

Sim-exp

Exp

Sim

Sim-exp

(1) (2) (3) (4) (5) (6)

80.4 70.3 72.8 66.4 71.5 76.5

80.7 71.5 73.9 65.8 71.0 73.8

0.3 1.2 1.1 −0.6 −0.5 −2.7

102.8 88.0 91.1 84.4 90.0 95.7

103.8 90.5 92.5 83.0 89.6 94

1.0 2.5 1.4 −1.4 −0.4 −1.7

120.3 105.1 108.2 101.1 107.5 112.7

121.1 108.0 109.0 102.0 108.2 111.7

0.8 2.9 0.8 0.9 0.7 −1.0

(a) Near the suction side

(b) On the opposite side of the suction side

Fig. 10. Temperature distribution outside the oil sump under GX conditions.

Fig. 11. Longitudinal section view of temperature distribution in the oil sump under GX conditions.

the bottom of cylinder and then increased towards the bottom of the sub bearing. The reason are as follows: (1) most parts of the main bearing were in direct contact with the discharge gas and discharge gas in the muffler with a high velocity and temperature heats the main bearing; (2) most parts of the cylinder were in direct contact with the low-temperature suction gas and; (3) most parts of the sub bearing were in contact with the oil sump whose temperature is lower than that of the discharge gas but higher than that of the suction gas. In the case of the cylinder, Fig. 13(b)–(e) show that in the circumferential direction, cylinder temperature increased gradually from the suction side to the discharge side by about 10 ℃. Fig. 13(a)–(e) show that in the radial direction, cylinder temperature decreased from outer surface to inner surface by more than 3 ℃. This is because oil temperature is higher than refrigerant temperature in the suction chamber and it transfers heat to the refrigerant. Near the discharge side, cylinder temperature was uniform as there was less temperature difference between the temperature of the refrigerant in the discharge chamber and oil temperature. Fig. 13(b)–(e) also show that in the vertical direction, cylinder temperature decreased from the top to the

respect to the middle of the sliding vane, respectively. Fig. 12 shows the cross-sectional view of the oil sump at heights of 55, 45, 40, 35, 20, 15, and 5 mm from the bottom of compressor shell. From Fig. 11(a) and (b) and Fig. 12 (a)–(e), it is found that oil temperature at the top of the sump was 86 ℃ and reduced to 70 ℃ at the bottom. This decrease can be attributed to heat dissipation by the oil to the compression unit and compressor shell while flowing downward. Oil temperature decreased by a large extent in the short distance from the top of the sump to the middle (about 12 ℃) after which it reduced slowly. This is because the low-temperature suction gas in the suction tube and suction chamber cools the oil. 4.2. The analysis of temperature distribution of compression unit Temperature-distribution contours of the compression unit under GX conditions are shown in Fig. 13. Fig. 13(c)–(e) show the cross-sectional views of the compression unit at heights of 45, 40, and 35 mm from the bottom of the compressor, respectively. In the vertical direction, compression unit temperature decreased from the main bearing to 8

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Fig. 12. Cross-sectional view of temperature contours in the oil sump under GX conditions. (a) 1–1 section view, (b) 2–2 section view, (c) 3–3 section view, (d) 4–4 section view, (e) 5–5 section view, (f) 6–6 section view, (g) 7–7 section view.

(a) longitudinal section view cut from the middle of the sliding vane (b) inner wall of the cylinder

(c) cross section view 1-1 (d) cross section view 2-2

(e) cross section view 3-3

Fig. 13. Temperature-distribution contours of the compression unit under GX conditions.

bottom, and the inner wall temperature has reduced by 4 ℃. Meanwhile, close to the discharge side, the temperature was stable and the reason for this uniformity is similar to that in the radial direction. Fig. 14 shows temperature distribution in the cylinder in cross section view 2–2 under different conditions. Variations in temperature under different conditions in the same direction were similar while the temperature values were different under different conditions. Under any condition, the highest temperature was close to that on the right side of the sliding vane and the lowest temperature was close to that on the left-side of the suction port. It can be seen from Fig. 14 that

temperature at a given point in the cylinder was the lowest under GX conditions, followed by ARI, and the highest is under ASHRAE/T1 condition.

4.3. The heat exchange between different parts Heat exchange between different parts is constant under steadystate conditions. Heat exchange between different parts and the ratio to indicated power are tabulated in Table 6 and the notations ①–⑫ are defined in Fig. 15. The indicated power can be calculated with the 9

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(a) GX condition

(b) ARI condition

(c) ASHRAE/T1 condition

Fig. 14. Temperature distribution in cross section view 2–2 under different conditions.

Table 6 Heat exchange between different parts of the oil sump and compression unit (W). Heat flow direction

①→② ⑤→⑩ ⑧→③ ⑧→① ③→④ ⑩⑫ → ④ ⑦→⑥ ②→⑨ ①→⑪

GX condition

ARI condition

ASHRAE/T1 condition

Heat (W)

Pi (W)

Ratio (%)

Heat (W)

Pi (W)

Ratio (%)

Heat (W)

Pi (W)

Ratio (%)

24.93 44.45 16.93 0.14 35.46 26.05 15.29 10.56 10.34

754.7 754.7 754.7 754.7 754.7 754.7 754.7 754.7 754.7

3.3 5.9 2.2 0.0 4.7 3.5 2.0 1.4 1.4

46.67 48.99 17.39 0.33 46.88 35.93 17.44 13.12 24.33

916.4 916.4 916.4 916.4 916.4 916.4 916.4 916.4 916.4

5.1 5.3 1.9 0.0 5.1 3.9 1.9 1.4 2.7

47.81 50.28 16.01 0.34 43.81 31.64 20.87 18.12 33.66

906.3 906.3 906.3 906.3 906.3 906.3 906.3 906.3 906.3

5.3 5.5 1.8 0.0 4.8 3.5 2.3 2.0 3.7

Note: (Ratio = 100 × Absolute Heat/Indicated power (Pi); →represents the heat-flow direction)

Fig. 15. Representation of the notations ①–⑫ in Table 5. Fig. 16. Heat dissipated from the oil sump.

known input power, the known motor efficiency (84%) and the mechanical efficiency (being about 94% within the three conditions, obtained from the dynamic and thermodynamic simulation model of rotary compressor [30]). Oil sump: In Fig. 16, it can be observed that the heat dissipated from the oil sump was different under different conditions; of this heat, a large proportion (about 65%) was transferred to the compression unit (main bearing, sub bearing, and cylinder) and the remaining (about 35%) was transferred to the compressor shell. While, the oil sump absorbs little heat from the sliding vane when compared to the compression unit and compressor shell (described in Table 6). It is worth noting that a large part of heat from the oil sump could heat the suction

gas through the compression unit, which may lead to performance and reliability problems. Compression unit: Table 6 shows that both the discharge gas and oil sump heated the compression unit, while the gas in the working chamber, suction tube, and compressor shell absorbed heat from the compression unit. Fig. 17 shows that the heat transferred to the compression unit by the discharge gas (accounts for about 5.5% of the indicated power (Pi)) and sliding vane (accounts for about 2% of Pi) varied little under different conditions but the heat transferred by the oil sump was different under different conditions. Among the studied conditions, heat absorbed by the compression unit (main bearing) from 10

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constant. Table 8 indicates that among the studied conditions, the refrigerant in the suction tube and working chamber absorbed about 10% of the indicated power (Pi) from the oil sump and discharge gas. Further, the relative errors of total-1 and total-2 show that the compression unit could maintain heat balance with the oil sump and refrigerant in the allowed error margin, which also agrees with the law of energy conservation. 4.5. Improvement It’s known from above analysis that about 10% of indicated power was absorbed by the refrigerant in the working chamber, which has a negative effect on the performance and reliability of R32 rotary compressor. Preventing this heat from heating the compression unit could decrease the superheat of gas in the suction chamber. A method of adding a layer of adiabatic coating on the outer surface of compression unit is proposed in this study, besides, in order to cool the oil temperature and keep the properties of oil in the outlet of oil sump, the outside of compressor shell was simultaneously ventilated (2 m/s). It is shown in Fig. 18 that after adding the adiabatic coating the oil temperature in the outlet of oil sump is approximately the same with the data before adding the adiabatic coating. As expected, the temperature of cylinder especially at the suction side decreased by more than 4 ℃, which would bring in the lower superheat of the gas in the suction chamber. The cooling capacity would be improved and the discharge gas temperature would be decreased, and then the performance and reliability of compressor could be improved.

Fig. 17. Heat transferred to the compression unit. Table 7 The relationship between enthalpy variation ΔH and total dissipation heat Q. Condition

ΔH (W)

Q (W)

Absolute error (W)

Relative error (%)

GX ARI ASHRAE/T1

34.0 70.3 80.2

33.3 68.8 82.0

0.7 1.5 1.8

2.1 2.1 2.2

5. Conclusions

the discharge gas was the highest, followed by that absorbed from the oil sump and sliding vane.

In this study, the 3-D CFD fluid-solid coupled method were proposed to analyze the temperature distribution and heat transfer mechanism of oil sump and compression unit in a R32 hermetic rotary compressor. The accuracy (less than 3 ℃) of the model is verified by the experimental results. The main conclusions in this study are as follows: (1) In the vertical direction of oil sump, the oil temperature drops by about 15 ℃ from top to bottom of oil sump, and decreases greatly above the middle of oil sump (about 12 ℃), and mildly below the middle of oil sump. The oil sump temperature near the outlet of oil sump is a little higher than that in other place of the bottom of oil sump. (2) In the circumferential direction, cylinder temperature increased by about 10 ℃ from the suction side to discharge side. At the same time, it decreased by about 4 ℃ from the top to the bottom and by more than 3 ℃ from outside the cylinder to its interior. Near the discharge port, the cylinder temperature was almost uniform. (3) Heat dissipated from the oil sump was different under different conditions, of which a large proportion (about 65%) was transferred to the compression unit and the remaining (about 35%) was transferred to the compressor shell. Furthermore, the oil sump absorbed little heat from the sliding vane when compared to the compression unit and

4.4. The heat balance calculation 1) Oil sump It is found from Table 6 that the oil sump dissipated heat to the compression unit and compressor shell. The total heat (Q) dissipated by the oil sump under different conditions is shown in Table 7. Enthalpy variation (ΔH) of oil between inlets (inlet I and inlet II) and the outlet was calculated using the following equation.

ΔH = CpinletI ·ṁ inletI ·(Tout − TinletI ) + CpinletII ·ṁ inletI ·(Tout − TinletII )

(22)

The values of ΔH under the three tested conditions are shown in Table 7 and it shows that within the allowed error margin, the total heat dissipated, Q, was equal to the enthalpy variation, ΔH, which agrees with the law of energy conservation. The errors may come from the solution method and the residual set in the solver. 2) Compression unit When a compressor works steadily, heat exchange between the compression unit and oil sump, refrigerant, and ambient environment is

Table 8 Heat exchange between the compression unit and oil sump, refrigerant, and ambient environment. Conditions

GX condition

ARI condition

ASHRAE/T1 condition

Item

Heat flow direction

Heat (W)

Ratio with Pi (%)

Heat (W)

Ratio with Pi (%)

Heat (W)

Ratio with Pi (%)

Heat absorbed by compression unit (W)

①→② ⑤→⑩ ⑧→③ ⑨→② Total-1 ⑩⑫ → ④ ⑦→⑥ ③→④ Total-2 (−)

24.93 44.45 16.93 −10.56 75.75 26.05 15.29 35.46 76.84 −1.39

3.3 5.9 2.2 −1.4 10.0 3.5 2.0 4.7 10.2 −1.39

46.67 48.99 17.39 −13.12 99.93 35.93 17.44 46.88 100.25 −0.32

5.1 5.3 1.9 −1.4 10.9 3.9 1.9 5.1 10.9 −0.32

47.81 50.28 16.01 −18.12 95.98 31.64 20.87 43.81 96.32 −0.35

5.3 5.5 1.8 −2.0 10.6 3.5 2.3 4.8 10.6 −0.35

Heat dissipated to working gas from compression unit (W)

Relative error (%) (Total-1 − Total-2)/Total-1

11

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Fig. 18. Comparisons between improved results and original results under different conditions.

the discharge gas by the compression unit (main bearing) was the highest followed by heat from the oil sump and sliding vane. (5) As heat absorbed by the refrigerant in the working chamber and suction tube accounts for about 10% of the indicated power, a method that adds an adiabatic coating on the surface of compression unit and simultaneously ventilates the outside of the compressor shell was

compressor shell. (4) Heat transferred to the compression unit from the discharge gas (accounts for about 5.5% of the indicated power) and sliding vane (accounts for about 2% of the indicated power) both varied little under different conditions, but heat from the oil sump was different under different conditions. Among the studied conditions, heat absorbed from 12

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proposed. Through this method, the temperature of cylinder especially at the suction side decreased by more than 4 ℃ under the studied conditions.

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